Method for determining a dilution of recirculated gases in a split exhaust engine

ABSTRACT

Methods and systems are provided for determining a dilution of recirculated gases, including blowthrough air, combusted exhaust gas, and fuel vapor, in a split exhaust engine. In one example, the dilution rate may be calculated using a steady state model based on temperature measurements from an intake region of the engine. The steady state model may be further used in combination with a feedforward model adapted to estimate engine dilution during engine transients as a correction factor.

FIELD

The present description relates generally to methods and systems fordetermining a dilution value or rate of recirculated gases in a splitexhaust engine operating with blowthrough.

BACKGROUND/SUMMARY

Engines may use boosting devices, such as turbochargers, to increaseengine power density. However, engine knock may occur due to increasedcombustion temperatures. Knock is especially problematic under boostedconditions due to high charge temperatures. The inventors herein haverecognized that utilizing an engine system with a split exhaust system,where a first exhaust manifold routes exhaust gas recirculation (EGR) toan intake of the engine, upstream of a compressor of the turbocharger,and where a second exhaust manifold routes exhaust to a turbine of theturbocharger in an exhaust of the engine, may decrease knock andincrease engine efficiency. In such an engine system, each cylinder mayinclude two intake valves and two exhaust valves, where a first set ofcylinder exhaust valves (e.g., blowdown exhaust valves) exclusivelycoupled to the first exhaust manifold may be operated at a differenttiming than a second set of cylinder exhaust valves (e.g., scavengeexhaust valves) exclusively coupled to the second exhaust manifold,thereby isolating a blowdown portion and scavenging portion of exhaustgases. The timing of the second set of cylinder exhaust valves may alsobe coordinated with a timing of cylinder intake valves to create apositive valve overlap period where fresh intake air (or a mixture offresh intake air and EGR), referred to as blowthrough, may flow throughthe cylinders and back to the intake, upstream of the compressor, via anEGR passage coupled to the first exhaust manifold. Blowthrough air mayremove residual exhaust gases from within the cylinders (referred to asscavenging). The inventors herein have recognized that by flowing afirst portion of the exhaust gas (e.g., higher pressure exhaust) throughthe turbine and a higher pressure exhaust passage and flowing a secondportion of the exhaust gas (e.g., lower pressure exhaust) andblowthrough air to the compressor inlet, combustion temperatures can bereduced while improving the turbine's work efficiency and engine torque.

However, the inventors herein have recognized potential issues with suchsystems. As one example, in the engine system described above, acomposition of gas recirculated to the intake may be more complex than atraditional EGR system comprising a single exhaust manifold or a systemthat does not recirculate increased volumes of blowthrough air. Whereasrecirculated gas in traditional EGR systems is entirely comprised ofburnt gas, the gas recirculated through the split exhaust engine mayinclude varying portions of burnt gas, fresh air, and pushback (e.g.,unburnt or non-combusted) fuel. Timing adjustment of engine operationssuch as fuel injection, spark advance, and intake and exhaust valveactuation timings based on an assumed EGR gas composition of traditionalEGR systems may result in degraded engine performance in the splitexhaust engine. Thus, a method to determine the composition of therecirculated gas in the split exhaust engine including blowthrough basedon a unique configuration of the engine is desirable to estimate an EGRdilution rate (e.g., a dilution value or rate of the gases recirculatedto the intake passage) for accurate engine control.

In one example, the issues described above may be addressed by a methodfor determining a dilution rate of gas recirculated from a first set ofexhaust valves to an intake passage via a recirculation passage based ona temperature of gases in each of the recirculation passage and theintake passage, upstream and downstream of where the EGR passage couplesto the intake passage, while flowing exhaust gas from a second set ofexhaust valves to a turbocharger turbine and not to the intake passage.As one example, a period of delay may occur where the recirculated gasmixture may be travelling through the engine before arriving at thecylinders for combustion. During this period, engine dilution may beapproximated by a steady state model based on the temperaturemeasurements at regions upstream and downstream of the merging point ofthe EGR passage to the intake passage.

In addition, the steady state model may be used to correct a feedforwardmodel to gain a more accurate estimate of the EGR rate to furtherimprove engine performance. The feedforward model evaluates a totalexhaust gas recirculation (EGR) mass flow, a temperature of the EGR gas,a mass of burnt gas, an airmass due to blowthrough, a fuel mass due toblowthrough, and a burnt gas fraction during engine transient events isused to estimate the EGR dilution rate. By accounting for thetemperature (i.e. heat capacity) contributions of each component of thegas mixture (including fresh air and recirculated gas, or scavenge gas)circulated through the split exhaust engine and an effect of a pressuredifferential across an intake region, an estimation of the EGR rate thatis corrected based on measured gas temperatures, may be tailored to anarchitecture of the split exhaust engine. As a result, engine operationssuch as fuel injection and spark timing may be adjusted to increase anefficiency and power output of the engine.

It should be understood that the summary above is provided to introducein simplified form a selection of concepts that are further described inthe detailed description. It is not meant to identify key or essentialfeatures of the claimed subject matter, the scope of which is defineduniquely by the claims that follow the detailed description.Furthermore, the claimed subject matter is not limited toimplementations that solve any disadvantages noted above or in any partof this disclosure.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 shows a schematic depiction of a turbocharged engine system witha split exhaust system.

FIG. 2 shows an embodiment of a cylinder of the engine system of FIG. 1.

FIG. 3 shows example cylinder intake valve and exhaust valve timings forone engine cylinder of a split exhaust engine system.

FIG. 4 shows a flow diagram of the variables used to calculate an EGRdilution rate flowing through a split exhaust system.

FIG. 5 shows an example map of average scavenge valve runner temperatureas a function of brake mean effective pressure for calculating atemperature of EGR in a scavenge manifold.

FIG. 6A shows an example map of pressure as function of crank angle atlower EGR flow, used for calculating a total EGR mass flow rate.

FIG. 6B shows an example map of pressure as function of crank angle athigher EGR flow, used for calculating a total EGR mass flow rate.

FIG. 7 shows an example routine for estimating a dilution rate ofrecirculated gases in a split exhaust engine.

FIG. 8 shows example adjustments to engine operating parametersresponsive to a determined dilution rate of gases recirculated to anintake in an intake of a split exhaust engine.

DETAILED DESCRIPTION

The following description relates to systems and methods for operating asplit exhaust engine with blowthrough and exhaust gas recirculation(EGR) to an intake via a scavenge exhaust manifold. As shown in FIG. 1,the split exhaust engine includes a first exhaust manifold (referred toherein as a blowdown exhaust manifold) coupled exclusively to a blowdownexhaust valve of each cylinder. The blowdown manifold is coupled to anexhaust passage of the engine, where the exhaust passage includes aturbocharger turbine and one or more emission control devices (which mayinclude one or more catalysts). The split exhaust engine may alsoinclude a second exhaust manifold (referred to herein as a scavengeexhaust manifold) coupled exclusively to a scavenge exhaust valve ofeach cylinder. The scavenge manifold is coupled to the intake passage,upstream of a turbocharger compressor, via a first EGR passage includinga first EGR valve (referred to herein as a scavenge EGR valve). In someembodiments, the split exhaust engine system may include additionalpassages coupled between the scavenge manifold and either the intake orexhaust passage, as shown in FIG. 1. Additionally, in some embodiments,the split exhaust engine system may include various valve actuationmechanisms and may be installed in a hybrid vehicle, as shown in FIG. 2.The scavenge exhaust valves and blowdown exhaust valves open and closeat different times in an engine cycle, for each cylinder, in order toisolate scavenge and blowdown portions of combusted exhaust gases anddirect these portions separately to the scavenge manifold and blowdownmanifold. As shown in FIG. 3, an overlap period may exist between theintake valves and the scavenge exhaust valve of each cylinder wherethese valves are open at the same time. As a result, fresh, blowthroughair may flow into the EGR passage via the scavenge exhaust valve. Thus,during each engine cycle, the EGR passage may receive a combination ofcombusted exhaust gases, blowthrough air, and unburnt fuel andrecirculate these combined gases to the intake passage. The specificarchitecture of the split exhaust engine system (as shown in the exampleof FIG. 1) and the timings of the cylinder valves which results in anincreased proportion of blowthrough air in the gases recirculated to theintake via the EGR passage (compared to a traditional engine with EGR)makes it challenging to determine the dilution rate of gasesrecirculated to the intake via the EGR passage and travelling to theengine cylinders. In some examples, this dilution rate of recirculatedgases may be referred to as the EGR rate; however, this EGR rate is theEGR rate of gases that include fresh, blowthrough air and unburnt fuel,in addition to combusted exhaust gas. In one example, the EGR rate(e.g., dilution rate of recirculated gases from the EGR passage to theintake passage) may be determined based on various engine operatingconditions and mapped parameters/models, as shown in FIG. 4.Specifically, FIG. 4 shows calculated variables contributing to afeedforward model of EGR rate based on mapped parameters and a steadystate correction which are used to estimate the EGR rate of the splitexhaust engine. As an example for calculating a temperature of EGR gasat a region where the scavenge manifold intersects with an intakepassage, a map of average scavenge runner temperature as a function ofbrake mean effective pressure is shown in FIG. 5. Example plots ofpressures detected at certain regions of the engine as a function ofengine crank angle are provided, illustrating a low EGR flow situationat FIG. 6A and a high EGR flow situation at FIG. 6B. The plots of FIGS.6A-6B may be used to calculate a total EGR mass flow rate. An exampleroutine for determining the EGR dilution rate and adjusting engineoperations in response to the dilution rate is shown in FIG. 7 withrespect to an engine controller. Example operations of a steady statemodel calculating a composition of gas flowing through the split exhaustengine used to correct the feedforward model are illustrated herein withreference to FIG. 8.

The model for EGR (e.g., dilution) rate determination shown in FIG. 4may represent an estimation of the EGR flow rate during enginetransients (e.g., transient engine operation). However, subsequent toperiods of low EGR flow, such as during low engine speeds and loads, anincrease in the EGR rate, as determined by the feedforward model, maydeviate from a true EGR rate at the intake manifold due to a delay intime for the circulated gas mixture to be ingested at the enginecylinders. Thus, a steady state model for approximating EGR flow may beused to adjust timing of engine operations until the mixture of scavengegas and fresh air is delivered to the engine intake. The steady statemodel may be determined based on a calculated temperature gradientacross a region upstream of the intake manifold. In some examples, useof the steady state model may be continued even after the feedforwardmodel becomes more effective as a temperature-derived correction tomodel the estimated EGR rate towards the true rate. In such events, thesteady state model may be used as a correction for the feedforwardmodel. However, in other examples, the feedforward model may be usedsolely to determine the dilution rate. Determination of the EGR rateestimates according to variables incorporated into both the transient,feedforward model, and the steady state correction are described herein.

In addition, any temperature sensor that depends upon adopting atemperature of the measured fluid may have an inherent time constantbased on a surface to mass ratio. Decreasing a diameter of a sensingelement of the sensor to that of a whisker, for example, may reduce thetime constant to an insignificant value. However, manufacturing of suchsmall diameter sensing elements may not be cost or time efficient. Thus,temperature measurements based on conventional sensors may be slowrelative to how quickly the exhaust gas fraction in an air stream maychange. While the temperature-based method for determining exhaustconcentration in the engine air flow has increased accuracy, it may beslower to react to change, thus the feedforward method for fasterresponses to transient events, combined with the steady state flowmeasure for accuracy or adaptation, may increase the overall accuracy ofthe EGR rate estimates.

FIG. 1 shows a schematic diagram of a multi-cylinder internal combustionengine 10, which may be included in a propulsion system of anautomobile. Engine 10 includes a plurality of combustion chambers (i.e.,cylinders) which may be capped on the top by a cylinder head (notshown). In the example shown in FIG. 1, engine 10 includes cylinders 12,14, 16, and 18, arranged in an inline-4 configuration. It should beunderstood, however, that though FIG. 1 shows four cylinders, engine 10may include any number of cylinders in any configuration, e.g., V-6,1-6, V-12, opposed 4, etc. Further, the cylinders shown in FIG. 1 mayhave a cylinder configuration, such as the cylinder configuration shownin FIG. 2, as described further below.

Each of cylinders 12, 14, 16, and 18 include two intake valves,including first intake valve 2 and second intake valve 4, and twoexhaust valves, including first exhaust valve (referred to herein as ablowdown exhaust valve, or blowdown valve) 8 and second exhaust valve(referred to herein as a scavenge exhaust valve, or scavenge valve) 6.The intake valves and exhaust valves may be referred to herein ascylinder intake valves and cylinder exhaust valves, respectively. Asexplained further below with reference to FIG. 2, a timing (e.g.,opening timing, closing timing, opening duration, etc.) of each of theintake valves may be controlled via various camshaft timing systems. Inone embodiment, both the first intake valves 2 and second intake valves4 may be controlled to a same valve timing (e.g., such that they openand close at the same time in the engine cycle). In an alternateembodiment, the first intake valves 2 and second intake valves 4 may becontrolled at a different valve timing. Further, the first exhaustvalves 8 may be controlled at a different valve timing than the secondexhaust valves 6 (e.g., such that a first exhaust valve and secondexhaust valve of a same cylinder open at different times than oneanother and close at different times than one another), as discussedfurther below.

Each cylinder receives intake air (or a mixture of intake air andrecirculated exhaust gas, as explained further below) from an intakemanifold 44 via an air intake passage 28. Intake manifold 44 is coupledto the cylinders via intake ports (e.g., runners). For example, intakemanifold 44 is shown in FIG. 1 coupled to each first intake valve 2 ofeach cylinder via first intake ports 20. Further, the intake manifold 44is coupled to each second intake valve 4 of each cylinder via secondintake ports 22. In this way, each cylinder intake port can selectivelycommunicate with the cylinder it is coupled to via a corresponding oneof the first intake valves 2 or second intake valves 4. Each intake portmay supply air and/or fuel to the cylinder it is coupled to forcombustion.

One or more of the intake ports may include a charge motion controldevice, such as a charge motion control valve (CMCV). As shown in FIG.1, each first intake port 20 of each cylinder includes a CMCV 24. CMCVs24 may also be referred to as swirl control valves or tumble controlvalves. CMCVs 24 may restrict airflow entering the cylinders via firstintake valves 2. In the example of FIG. 1, each CMCV 24 may include avalve plate; however, other configurations of the valve are possible.Note that for the purposes of this disclosure the CMCV 24 is in the“closed” position when it is fully activated and the valve plate may befully tilted into the respective first intake port 20, thereby resultingin maximum air charge flow obstruction. Alternatively, the CMCV 24 is inthe “open” position when deactivated and the valve plate may be fullyrotated to lie substantially parallel with airflow, thereby considerablyminimizing or eliminating airflow charge obstruction. The CMCVs mayprincipally be maintained in their “open” position and may only beactivated “closed” when swirl conditions are desired. As shown in FIG.1, only one intake port of each cylinder includes the CMCV 24. However,in alternate embodiments, both intake ports of each cylinder may includea CMCV 24. The controller 12 may actuate the CMCVs 24 (e.g., via a valveactuator that may be coupled to a rotating shaft directly coupled toeach CMCV 24) to move the CMCVs into the open or closed positions, or aplurality of positions between the open and closed positions, inresponse to engine operating conditions (such as engine speed/loadand/or when blowthrough via the second exhaust valves 6 is active), asexplained further below. As referred to herein, blowthrough air orblowthrough combustion cooling may refer to intake air that flows fromthe one or more intake valves of each cylinder to second exhaust valves6 (and into second exhaust manifold 80) during a valve opening overlapperiod between the intake valves and second exhaust valves 6 (e.g., aperiod when both the intake valves and second exhaust valves 6 are openat the same time), without combusting the blowthrough air.

A high pressure, dual stage, fuel system (such as the fuel system shownin FIG. 2) may be used to generate fuel pressures at injectors 66. Assuch, fuel may be directly injected in the cylinders via injectors 66.Distributorless ignition system 88 provides an ignition spark tocylinders 12, 14, 16, and 18 via sparks plug 92 in response tocontroller 12. Cylinders 12, 14, 16, and 18 are each coupled to twoexhaust ports for channeling the blowdown and scavenging portions of thecombustion gases separately. Specifically, as shown in FIG. 1, cylinders12, 14, 16, and 18 exhaust combustion gases (e.g., scavenging portion)to second exhaust manifold (referred to herein as a scavenge manifold)80 via second exhaust runners (e.g., ports) 82 and combustion gases(e.g., blowdown portion) to first exhaust manifold (referred to hereinas a blowdown manifold) 84 via first exhaust runners (e.g., ports) 86.Second exhaust runners 82 extend from cylinders 12, 14, 16, and 18 tosecond exhaust manifold 80. Additionally, first exhaust manifold 84includes a first manifold portion 81 and second manifold portion 85.First exhaust runners 86 of cylinders 12 and 18 (referred to herein asthe outside cylinders) extend from cylinders 12 and 18 to the secondmanifold portion 85 of first exhaust manifold 84. Additionally, firstexhaust runners 86 of cylinders 14 and 16 (referred to herein as theinside cylinders) extend from cylinders 14 and 16 to the first manifoldportion 81 of first exhaust manifold 84.

Each exhaust runner can selectively communicate with the cylinder it iscoupled to via an exhaust valve. For example, second exhaust runners 82communicate with their respective cylinders via second exhaust valves 6and first exhaust runners 86 communicate with their respective cylindersvia first exhaust valves 8. Second exhaust runners 82 are isolated fromfirst exhaust runners 86 when at least one exhaust valve of eachcylinder is in a closed position. Exhaust gases may not flow directlybetween exhaust runners 82 and 86. The exhaust system described abovemay be referred to herein as a split exhaust manifold system, where afirst portion of exhaust gases from each cylinder are output to firstexhaust manifold 84 and a second portion of exhaust gases from eachcylinder are output to second exhaust manifold 80, and where the firstand second exhaust manifolds do not directly communicate with oneanother (e.g., no passage directly couples the two exhaust manifolds toone another and thus the first and second portions of exhaust gases donot mix with one another within the first and second exhaust manifolds).

Engine 10 includes a turbocharger including a dual-stage exhaust turbine164 and an intake compressor 162 coupled on a common shaft. Dual-stageturbine 164 includes a first turbine 163 and second turbine 165. Firstturbine 163 is directly coupled to first manifold portion 81 of firstexhaust manifold 84 and receives exhaust gases only from cylinders 14and 16 via first exhaust valves 8 of cylinders 14 and 16. Second turbine165 is directly coupled to second manifold portion 85 of first exhaustmanifold 84 and receives exhaust gases only from cylinders 12 and 18 viafirst exhaust valves 8 of cylinders 12 and 18. Rotation of first andsecond turbines drives rotation of compressor 162 disposed within theintake passage 28. As such, the intake air becomes boosted (e.g.,pressurized) at the compressor 162 and travels downstream to intakemanifold 44. Exhaust gases exit both first turbine 163 and secondturbine 165 into common exhaust passage 74. A wastegate may be coupledacross the dual-stage turbine 164. Specifically, wastegate valve 76 maybe included in a bypass 78 coupled between each of the first manifoldportion 81 and second manifold portion 85, upstream of an inlet todual-stage turbine 164, and exhaust passage 74, downstream of an outletof dual-stage turbine 164. In this way, a position of wastegate valve(referred to herein as a turbine wastegate) 76 controls an amount ofboost provided by the turbocharger. In alternate embodiments, engine 10may include a single stage turbine where all exhaust gases from thefirst exhaust manifold 84 are directed to an inlet of a same turbine.

Exhaust gases exiting dual-stage turbine 164 flow downstream in exhaustpassage 74 to a first emission control device 70 and a second emissioncontrol device 72, second emission control device 72 arranged downstreamin exhaust passage 74 from first emission control device 70. Emissioncontrol devices 70 and 72 may include one or more catalyst bricks, inone example. In some examples, emission control devices 70 and 72 may bethree-way type catalysts. In other examples, emission control devices 70and 72 may include one or a plurality of a diesel oxidation catalyst(DOC), and a selective catalytic reduction catalyst (SCR). In yetanother example, second emission control device 72 may include agasoline particulate filter (GPF). In one example, first emissioncontrol device 70 may include a catalyst and second emission controldevice 72 may include a GPF. After passing through emission controldevices 70 and 72, exhaust gases may be directed out to a tailpipe.

Exhaust passage 74 further includes a plurality of exhaust sensors inelectronic communication with controller 12 of control system 15, asdescribed further below. As shown in FIG. 1, exhaust passage 74 includesa first oxygen sensor 90 positioned between first emission controldevice 70 and second emission control device 72. First oxygen sensor 90may be configured to measure an oxygen content of exhaust gas enteringsecond emission control device 72. Exhaust passage 74 may include one ormore additional oxygen sensors positioned along exhaust passage 74, suchas second oxygen sensor 91 positioned between dual-stage turbine 164 andfirst emission control device 70 and/or third oxygen sensor 93positioned downstream of second emission control device 72. As such,second oxygen sensor 91 may be configured to measure the oxygen contentof the exhaust gas entering first emission control device 70 and thirdoxygen sensor 93 may be configured to measure the oxygen content ofexhaust gas exiting second emission control device 72. In oneembodiment, the one or more oxygen sensor 90, 91, and 93 may beUniversal Exhaust Gas Oxygen (UEGO) sensors. Alternatively, a two-stateexhaust gas oxygen sensor may be substituted for oxygen sensors 90, 91,and 93. Exhaust passage 74 may include various other sensors, such asone or more temperature and/or pressure sensors. For example, as shownin FIG. 1, a pressure sensor 96 is positioned within exhaust passage 74,between first emission control device 70 and second emission controldevice 72. As such, pressure sensor 96 may be configured to measure thepressure of exhaust gas entering second emission control device 72. Bothpressure sensor 96 and oxygen sensor 90 are arranged within exhaustpassage 74 at a point where a flow passage 98 couples to exhaust passage74. Flow passage 98 may be referred to herein as a scavenge manifoldbypass passage (SMBP) 98. Scavenge manifold bypass passage 98 isdirectly coupled to and between second exhaust (e.g., scavenge) manifold80 and exhaust passage 74. A valve 97 (referred to herein as thescavenge manifold bypass valve, SMBV) is disposed within scavengemanifold bypass passage 98 and is actuatable by controller 12 to adjustan amount of exhaust flow from second exhaust manifold 80 to exhaustpassage 74, at a location between first emission control device 70 andsecond emission control device 72.

Second exhaust manifold 80 is directly coupled to a first exhaust gasrecirculation (EGR) passage 50. First EGR passage 50 is a coupleddirectly between second exhaust manifold 80 and intake passage 28,upstream of compressor (e.g., turbocharger compressor) 162 (and thus maybe referred to as a low-pressure EGR passage). As such, exhaust gases(or blowthrough air, as explained further below) is directed from secondexhaust manifold 80 to intake passage 28, upstream of compressor 162,via first EGR passage 50. First EGR passage 50 is shown in FIG. 1without an EGR cooler but in alternate embodiments, an EGR cooler may bearranged in first EGR passage 50 to cool exhaust gases flowing fromsecond exhaust manifold 80 to intake passage 28 and a first EGR valve 54(which may be referred to herein a scavenge EGR valve 54). Controller 12is configured to actuate and adjust a position of first EGR valve 54 inorder to control an amount of air flow through first EGR passage 50.When first EGR valve 54 is in a closed position, no exhaust gases orintake air may flow from second exhaust manifold 80 to intake passage28, upstream of compressor 162. Further, when first EGR valve 54 is inan open position, exhaust gases and/or blowthrough air may flow fromsecond exhaust manifold 80 to intake passage 28, upstream of compressor162. Controller 12 may additionally adjust first EGR valve 54 into aplurality of positions between fully open and fully closed.

A first ejector 56 is positioned at an outlet of EGR passage 50, withinintake passage 28. First ejector 56 may include a constriction orventuri that provides a pressure increase at the inlet of the compressor162. As a result, EGR from the EGR passage 50 may be mixed with freshair flowing through the intake passage 28 to the compressor 162. Thus,EGR from the EGR passage 50 may act as the motive flow on the firstejector 56. In an alternate embodiment, there may not be an ejectorpositioned at the outlet of EGR passage 50. Instead, an outlet ofcompressor 162 may be shaped as an ejector that lowers the gas pressureto assist in EGR flow (and thus, in this embodiment, air is the motiveflow and EGR is the secondary flow). In yet another embodiment, EGR fromEGR passage 50 may be introduced at the trailing edge of a blade ofcompressor 162, thereby allowing blowthrough air to intake passage 28via EGR passage 50.

A second EGR passage 58 is coupled between first EGR passage 50 andintake passage 28. Specifically, as shown in FIG. 1, second EGR passage58 is coupled to first EGR passage 50, upstream of EGR valve 54.Additionally, second EGR passage 58 is directly coupled to intakepassage 28, downstream of compressor 162. Due to this coupling, secondEGR passage 58 may be referred to herein as a mid-pressure EGR passage.Further, as shown in FIG. 1, second EGR passage 58 is coupled to intakepassage 28 upstream of a charge air cooler (CAC) 40. CAC 40 isconfigured to cool intake air (which may be a mixture of fresh intakeair from outside of the engine system and exhaust gases) as it passesthrough CAC 40. As such, recirculated exhaust gases from first EGRpassage 50 and/or second EGR passage 58 may be cooled via CAC 40 beforeentering intake manifold 44. In an alternate embodiment, second

EGR passage 58 may be coupled to intake passage 28, downstream of CAC40. Further, as shown in FIG. 1, a second ejector 57 may be positionedwithin intake passage 28, at an outlet of second EGR passage 58.

A second EGR valve 59 (e.g., mid-pressure EGR valve) is disposed withinsecond EGR passage 58. Second EGR valve 59 is configured to adjust anamount of gas flow (e.g., intake air or exhaust) through second EGRpassage 58. As described further below, controller 12 may actuate EGRvalve 59 into an open position (allowing flow thorough second EGRpassage 58), closed position (blocking flow through second EGR passage58), or plurality of positions between fully open and fully closed basedon (e.g., as a function of) engine operating conditions. For example,actuating the EGR valve 59 may include the controller 12 sending anelectronic signal to an actuator of the EGR valve 59 to move a valveplate of EGR valve 59 into an open position, closed position, or someposition between fully open and fully closed. As also explained furtherbelow, based on system pressures and positons of alternate valves in theengine system, air may either flow toward intake passage 28 withinsecond EGR passage 58 or toward second exhaust manifold 80 within secondEGR passage 58.

Intake passage 28 further includes an electronic intake throttle 62 incommunication with intake manifold 44. As shown in FIG. 1, intakethrottle 62 is positioned downstream of CAC 40. The position of athrottle plate 64 of throttle 62 can be adjusted by control system 15via a throttle actuator (not shown) communicatively coupled tocontroller 12. By modulating air intake throttle 62, while operatingcompressor 162, an amount of fresh air may be inducted from theatmosphere and/or an amount of recirculated exhaust gas from the one ormore EGR passages into engine 10, cooled by CAC 40 and delivered to theengine cylinders at compressor (or boosted) pressure via intake manifold44. To reduce compressor surge, at least a portion of the airchargecompressed by compressor 162 may be recirculated to the compressorinlet. A compressor recirculation passage 41 may be provided forrecirculating compressed air from the compressor outlet, upstream of CAC40, to the compressor inlet. Compressor recirculation valve (CRV) 42 maybe provided for adjusting an amount of recirculation flow recirculatedto the compressor inlet. In one example, CRV 42 may be actuated open viaa command from controller 12 in response to actual or expectedcompressor surge conditions.

A third flow passage 30 (which may be referred to herein as a hot pipe)is coupled between second exhaust manifold 80 and intake passage 28.Specifically, a first end of third flow passage 30 is directly coupledto second exhaust manifold 80 and a second end of third flow passage 30is directly coupled to intake passage 28, downstream of intake throttle62 and upstream of intake manifold 44. A third valve 32 (e.g., hot pipevalve) is disposed within third flow passage 30 and is configured toadjust an amount of air flow through third flow passage 30. Third valve32 may be actuated into a fully open position, fully closed position, ora plurality of positions between fully open and fully closed in responseto an actuation signal sent to an actuator of third valve 32 fromcontroller 12.

Second exhaust manifold 80 and/or second exhaust runners 82 may includeone or more sensors (such as pressure, oxygen, and/or temperaturesensors) disposed therein. For example, as shown in FIG. 1, secondexhaust manifold 80 includes pressure sensors 34 and 53, temperaturesensor 52, and oxygen sensor 36, disposed therein and configured tomeasure a pressure, a temperature, and an oxygen content, respectively,of exhaust gases and blowthrough (e.g., intake) air, exiting secondexhaust valves 6 and entering second exhaust manifold 80. Additionallyor alternatively to oxygen sensor 36, each second exhaust runner 82 mayinclude an individual oxygen sensor 38 disposed therein. As such, anoxygen content of exhaust gases and/or blowthrough air exiting eachcylinder via second exhaust valves 6 may be determined based on anoutput of oxygen sensors 38.

In some embodiments, as shown in FIG. 1, intake passage 28 may includean electric compressor 60. Electric compressor 60 is disposed in abypass passage 61 which is coupled to intake passage 28, upstream anddownstream of an electric compressor valve 63. Specifically, an inlet tobypass passage 61 is coupled to intake passage 28 upstream of electriccompressor valve 63 and an outlet to bypass passage 61 is coupled tointake passage 28 downstream of electric compressor valve 63 andupstream of where first EGR passage 50 couples to intake passage 28.Further, the outlet of bypass passage 61 is coupled upstream in intakepassage 28 from turbocharger compressor 162. Electric compressor 60 maybe electrically driven by an electric motor using energy stored at anenergy storage device. In one example, the electric motor may be part ofelectric compressor 60, as shown in FIG. 1. When additional boost (e.g.,increased pressure of the intake air above atmospheric pressure) isrequested, over an amount provided by compressor 162, controller 12 mayactivate electric compressor 60 such that it rotates and increases apressure of intake air flowing through bypass passage 61. Further,controller 12 may actuate electric compressor valve 63 into a closed orpartially closed position to direct an increased amount of intake airthrough bypass passage 61 and electric compressor 60.

Intake passage 28 may include one or more additional sensors (such asadditional pressure, temperature, flow rate, and/or oxygen sensors). Forexample, as shown in FIG. 1, intake passage 28 includes a mass air flow(MAF) sensor 48 and a first intake temperature sensor 3, disposedupstream of compressor 162, electric compressor valve 63, and wherefirst EGR passage 50 couples to intake passage 28. A first intakepressure sensor 51 may be arranged immediately upstream of the venturiof ejector 56. A second intake pressure sensor 31 and a second intaketemperature sensor 33 are positioned in intake passage 28, upstream ofcompressor 162 and downstream of where first EGR passage 50 couples tointake passage 28. An intake oxygen sensor 35 and an intake temperaturesensor 43 may be located in intake passage 28, downstream of compressor162 and upstream of CAC 40. An additional intake pressure sensor 37 maybe positioned in intake passage 28, downstream of CAC 40 and upstream ofthrottle 62. In some embodiments, as shown in FIG. 1, an additionalintake oxygen sensor 39 may be positioned in intake passage 28, betweenCAC 40 and throttle 62. Further, an intake manifold pressure (e.g., MAP)sensor 122 and intake manifold temperature sensor 123 are positionedwithin intake manifold 44, upstream of all engine cylinders.

In some examples, engine 10 may be coupled to an electric motor/batterysystem (as shown in FIG. 2) in a hybrid vehicle. The hybrid vehicle mayhave a parallel configuration, series configuration, or variation orcombinations thereof. Further, in some embodiments, other engineconfigurations may be employed, for example a diesel engine.

Engine 10 may be controlled at least partially by a control system 15including controller 12 and by input from a vehicle operator via aninput device (not shown in FIG. 1). Control system 15 is shown receivinginformation from a plurality of sensors 16 (various examples of whichare described herein) and sending control signals to a plurality ofactuators 83. As one example, sensors 16 may include pressure,temperature, and oxygen sensors located within the intake passage 28,intake manifold 44, exhaust passage 74, and second exhaust manifold 80,as described above. Other sensors may include a throttle inlet pressure(TIP) sensor for estimating a throttle inlet pressure (TIP) and/or athrottle inlet temperature sensor for estimating a throttle airtemperature (TCT) coupled downstream of the throttle in the intakepassage. Additional system sensors and actuators are elaborated belowwith reference to FIG. 2. As another example, actuators 83 may includefuel injectors, valves 63, 42, 54, 59, 32, 97, 76, and throttle 62.Actuators 83 may further include various camshaft timing actuatorscoupled to the cylinder intake and exhaust valves (as described furtherbelow with reference to FIG. 2). Controller 12 may receive input datafrom the various sensors, process the input data, and trigger theactuators in response to the processed input data based on instructionor code programmed in a memory of controller 12 corresponding to one ormore routines. Example control routines (e.g., methods) are describedherein at FIG. 7. For example, adjusting EGR flow from second exhaustmanifold 80 to intake passage 28 may include adjusting an actuator offirst EGR valve 54 to adjust an amount of exhaust flow flowing to intakepassage 28, upstream of compressor 162, from second exhaust manifold 80.In another example, adjusting EGR flow from second exhaust manifold 80to intake passage 28 may include adjusting an actuator of an exhaustvalve camshaft to adjust an opening timing of second exhaust valves 6.

In this way, the first and second exhaust manifolds of FIG. 1 may beconfigured to separately channel the blowdown and scavenging portions ofthe exhaust. First exhaust manifold 84 may channel the blowdown pulse ofthe exhaust to dual-stage turbine 164 via first manifold portion 81 andsecond manifold portion 85 while second exhaust manifold 80 may channelthe scavenging portion of exhaust to intake passage 28 via one or moreof first EGR passage 50 and second EGR passage 58 and/or to exhaustpassage 74, downstream of the dual-stage turbine 164, via flow passage98. For example, first exhaust valves 8 channel the blowdown portion ofthe exhaust gases through first exhaust manifold 84 to the dual-stageturbine 164 and both first and second emission control device 70 and 72while second exhaust valves 6 channel the scavenging portion of exhaustgases through second exhaust manifold 80 and to either intake passage 28via one or more EGR passages or exhaust passage 74 and second emissioncontrol device 72 via flow passage 98.

It should be noted that while FIG. 1 shows engine 10 including each offirst EGR passage 50, second EGR passage 58, flow passage 98, and flowpassage 30, in alternate embodiments, engine 10 may only include aportion of these passages. For example, in one embodiment, engine 10 mayonly include first EGR passage 50 and flow passage 98 and not includesecond EGR passage 58 and flow passage 30. In another embodiment, engine10 may include first EGR passage 50, second EGR passage 58, and flowpassage 98, but not include flow passage 30. In yet another embodiment,engine 10 may include first EGR passage 50, flow passage 30, and flowpassage 98, but not second EGR passage 58. In some embodiments, engine10 may not include electric compressor 60. In still other embodiments,engine 10 may include all or only a portion of the sensors shown in FIG.1.

Referring now to FIG. 2, it depicts a partial view of a single cylinderof internal combustion engine 10 which may be installed in a vehicle100. As such, components previously introduced in FIG. 1 are representedwith the same reference numbers and are not re-introduced. Engine 10 isdepicted with combustion chamber (cylinder) 130, coolant sleeve 114, andcylinder walls 132 with piston 136 positioned therein and connected tocrankshaft 140. Combustion chamber 130 is shown communicating withintake passage 146 and exhaust passage 148 via respective intake valve152 and exhaust valve 156. As previously described in FIG. 1, eachcylinder of engine 10 may exhaust combustion products along twoconduits. In the depicted view, exhaust passage 148 represents the firstexhaust runner (e.g., port) leading from the cylinder to the turbine(such as first exhaust runner 86 of FIG. 1) while the second exhaustrunner is not visible in this view.

As also previously elaborated in FIG. 1, each cylinder of engine 10 mayinclude two intake valves and two exhaust valves. In the depicted view,intake valve 152 and exhaust valve 156 are located at an upper region ofcombustion chamber 130. Intake valve 152 and exhaust valve 156 may becontrolled by controller 12 using respective cam actuation systemsincluding one or more cams. The cam actuation systems may utilize one ormore of cam profile switching (CPS), variable cam timing (VCT), variablevalve timing (VVT) and/or variable valve lift (VVL) systems to varyvalve operation. In the depicted example, each intake valve 152 iscontrolled by an intake cam 151 and each exhaust valve 156 is controlledby an exhaust cam 153. The intake cam 151 may be actuated via an intakevalve timing actuator 101 and the exhaust cam 153 may be actuated via anexhaust valve timing actuator 103 according to set intake and exhaustvalve timings, respectively. In some examples, the intake valves andexhaust valves may be deactivated via the intake valve timing actuator101 and exhaust valve timing actuator 103, respectively. For example,the controller may send a signal to the exhaust valve timing actuator103 to deactivate the exhaust valve 156 such that it remains closed anddoes not open at its set timing. The position of intake valve 152 andexhaust valve 156 may be determined by valve position sensors 155 and157, respectively. As introduced above, in one example, all exhaustvalves of every cylinder may be controlled on a same exhaust camshaft.As such, both a timing of the scavenge (second) exhaust valves and theblowdown (first) exhaust valves may be adjusted together via onecamshaft, but they may each have different timings relative to oneanother. In another example, the scavenge exhaust valve of everycylinder may be controlled on a first exhaust camshaft and a blowdownexhaust valve of every cylinder may be controlled on a different, secondexhaust camshaft. In this way, the valve timing of the scavenge valvesand blowdown valves may be adjusted separately from one another. Inalternate embodiments, the cam or valve timing system(s) of the scavengeand/or blowdown exhaust valves may employ a cam in cam system, anelectro-hydraulic type system on the scavenge valves, and/or anelectro-mechanical valve lift control on the scavenge valves.

For example, in some embodiments, the intake and/or exhaust valve may becontrolled by electric valve actuation. For example, cylinder 130 mayalternatively include an intake valve controlled via electric valveactuation and an exhaust valve controlled via cam actuation includingCPS and/or VCT systems. In still other embodiments, the intake andexhaust valves may be controlled by a common valve actuator or actuationsystem, or a variable valve timing actuator or actuation system.

In one example, intake cam 151 includes separate and different cam lobesthat provide different valve profiles (e.g., valve timing, valve lift,duration, etc.) for each of the two intake valves of combustion chamber130. Likewise, exhaust cam 153 may include separate and different camlobes that provide different valve profiles (e.g., valve timing, valvelift, duration, etc.) for each of the two exhaust valves of combustionchamber 130. In another example, intake cam 151 may include a commonlobe, or similar lobes, that provide a substantially similar valveprofile for each of the two intake valves.

In addition, different cam profiles for the different exhaust valves canbe used to separate exhaust gases exhausted at low cylinder pressurefrom exhaust gases exhausted at exhaust pressure. For example, a firstexhaust cam profile can open from closed position the first exhaustvalve (e.g., blowdown valve) just before BDC (bottom dead center) of thepower stroke of combustion chamber 130 and close the same exhaust valvewell before top dead center (TDC) to selectively exhaust blowdown gasesfrom the combustion chamber. Further, a second exhaust cam profile canbe positioned to open from close a second exhaust valve (e.g., scavengevalve) before a mid-point of the exhaust stroke and close it after TDCto selectively exhaust the scavenging portion of the exhaust gases.

Thus, the timing of the first exhaust valve and the second exhaust valvecan isolate cylinder blowdown gases from scavenging portion of exhaustgases while any residual exhaust gases in the clearance volume of thecylinder can be cleaned out with fresh intake air blowthrough duringpositive valve overlap between the intake valve and the scavenge exhaustvalves. By flowing a first portion of the exhaust gas leaving thecylinders (e.g., higher pressure exhaust) to the turbine(s) and a higherpressure exhaust passage and flowing a later, second portion of theexhaust gas (e.g., lower pressure exhaust) and blowthrough air to thecompressor inlet, the engine system efficiency is improved. Turbineenergy recovery may be enhanced and engine efficiency may be improvedvia increased EGR and reduced knock.

Continuing with FIG. 2, exhaust gas sensor 126 is shown coupled toexhaust passage 148. Sensor 126 may be positioned in the exhaust passageupstream of one or more emission control devices, such as devices 70 and72 of FIG. 1. Sensor 126 may be selected from among various suitablesensors for providing an indication of exhaust gas air/fuel ratio suchas a linear oxygen sensor or UEGO (universal or wide-range exhaust gasoxygen), a two-state oxygen sensor or EGO (as depicted), a HEGO (heatedEGO), a NOx, HC, or CO sensor, for example. The downstream emissioncontrol devices may include one or more of a three way catalyst (TWC),NOx trap, GPF, various other emission control devices, or combinationsthereof.

Exhaust temperature may be estimated by one or more temperature sensors(not shown) located in exhaust passage 148. Alternatively, exhausttemperature may be inferred based on engine operating conditions such asspeed, load, air-fuel ratio (AFR), spark retard, etc.

Cylinder 130 can have a compression ratio, which is the ratio of volumeswhen piston 136 is at bottom center to top center. Conventionally, thecompression ratio is in the range of 9:1 to 10:1. However, in someexamples where different fuels are used, the compression ratio may beincreased. This may happen, for example, when higher octane fuels orfuels with higher latent enthalpy of vaporization are used. Thecompression ratio may also be increased if direct injection is used dueto its effect on engine knock.

In some embodiments, each cylinder of engine 10 may include a spark plug92 for initiating combustion. Ignition system 188 can provide anignition spark to combustion chamber 130 via spark plug 92 in responseto spark advance signal SA from controller 12, under select operatingmodes. However, in some embodiments, spark plug 92 may be omitted, suchas where engine 10 may initiate combustion by auto-ignition or byinjection of fuel as may be the case with some diesel engines.

In some embodiments, each cylinder of engine 10 may be configured withone or more fuel injectors for providing fuel thereto. As a non-limitingexample, cylinder 130 is shown including one fuel injector 66. Fuelinjector 66 is shown coupled directly to combustion chamber 130 forinjecting fuel directly therein in proportion to the pulse width ofsignal FPW received from controller 12 via electronic driver 168. Inthis manner, fuel injector 66 provides what is known as direct injection(hereafter also referred to as “DI”) of fuel into combustion cylinder130. While FIG. 2 shows injector 66 as a side injector, it may also belocated overhead of the piston, such as near the position of spark plug92. Such a position may improve mixing and combustion when operating theengine with an alcohol-based fuel due to the lower volatility of somealcohol-based fuels. Alternatively, the injector may be located overheadand near the intake valve to improve mixing. In an alternate embodiment,injector 66 may be a port injector providing fuel into the intake portupstream of cylinder 130.

Fuel may be delivered to fuel injector 66 from a high pressure fuelsystem 180 including fuel tanks, fuel pumps, and a fuel rail.Alternatively, fuel may be delivered by a single stage fuel pump atlower pressure, in which case the timing of the direct fuel injectionmay be more limited during the compression stroke than if a highpressure fuel system is used. Further, while not shown, the fuel tanksmay have a pressure transducer providing a signal to controller 12. Fueltanks in fuel system 180 may hold fuel with different fuel qualities,such as different fuel compositions. These differences may includedifferent alcohol content, different octane, different heat ofvaporizations, different fuel blends, and/or combinations thereof etc.In some embodiments, fuel system 180 may be coupled to a fuel vaporrecovery system including a canister for storing refueling and diurnalfuel vapors. The fuel vapors may be purged from the canister to theengine cylinders during engine operation when purge conditions are met.For example, the purge vapors may be naturally aspirated into thecylinder via the first intake passage at or below barometric pressure.

Engine 10 may be controlled at least partially by controller 12 and byinput from a vehicle operator 113 via an input device 118 such as anaccelerator pedal 116. The input device 118 sends a pedal positionsignal to controller 12. Controller 12 is shown in FIG. 2 as amicrocomputer, including a microprocessor unit 102, input/output ports104, an electronic storage medium for executable programs andcalibration values shown as a read only memory 106 in this particularexample, random access memory 108, keep alive memory 110, and a databus. Storage medium read-only memory 106 can be programmed with computerreadable data representing instructions executable by microprocessor 102for performing the methods and routines described below as well as othervariants that are anticipated but not specifically listed. Controller 12may receive various signals from sensors coupled to engine 10, inaddition to those signals previously discussed, including measurement ofinducted mass air flow (MAF) from mass air flow sensor 48; enginecoolant temperature (ECT) from temperature sensor 112 coupled to coolantsleeve 114; a profile ignition pickup signal (PIP) from Hall effectsensor 120 (or other type) coupled to crankshaft 140; throttle position(TP) from a throttle position sensor; absolute manifold pressure signal(MAP) from sensor 122, cylinder AFR from EGO sensor 126, and abnormalcombustion from a knock sensor and a crankshaft acceleration sensor.Engine speed signal, RPM, may be generated by controller 12 from signalPIP. Manifold pressure signal MAP from a manifold pressure sensor may beused to provide an indication of vacuum, or pressure, in the intakemanifold.

Based on input from one or more of the above-mentioned sensors,controller 12 may adjust one or more actuators, such as fuel injector66, throttle 62, spark plug 92, intake/exhaust valves and cams, etc. Thecontroller may receive input data from the various sensors, process theinput data, and trigger the actuators in response to the processed inputdata based on instruction or code programmed therein corresponding toone or more routines.

In some examples, vehicle 100 may be a hybrid vehicle with multiplesources of torque available to one or more vehicle wheels 160. In otherexamples, vehicle 100 is a conventional vehicle with only an engine, oran electric vehicle with only electric machine(s). In the example shownin FIG. 2, vehicle 100 includes engine 10 and an electric machine 161.Electric machine 161 may be a motor or a motor/generator and thus mayalso be referred to herein as an electric motor. Crankshaft 140 ofengine 10 and electric machine 161 are connected via a transmission 167to vehicle wheels 160 when one or more clutches 166 are engaged. In thedepicted example, a first clutch 166 is provided between crankshaft 140and electric machine 161, and a second clutch 166 is provided betweenelectric machine 161 and transmission 167. Controller 12 may send asignal to an actuator of each clutch 166 to engage or disengage theclutch, so as to connect or disconnect crankshaft 140 from electricmachine 161 and the components connected thereto, and/or connect ordisconnect electric machine 161 from transmission 167 and the componentsconnected thereto. Transmission 167 may be a gearbox, a planetary gearsystem, or another type of transmission. The powertrain may beconfigured in various manners including as a parallel, a series, or aseries-parallel hybrid vehicle.

Electric machine 161 receives electrical power from a traction battery170 to provide torque to vehicle wheels 160. Electric machine 161 mayalso be operated as a generator to provide electrical power to chargebattery 170, for example during a braking operation.

FIGS. 1-2 show example configurations with relative positioning of thevarious components. If shown directly contacting each other, or directlycoupled, then such elements may be referred to as directly contacting ordirectly coupled, respectively, at least in one example. Similarly,elements shown contiguous or adjacent to one another may be contiguousor adjacent to each other, respectively, at least in one example. As anexample, components laying in face-sharing contact with each other maybe referred to as in face-sharing contact. As another example, elementspositioned apart from each other with only a space there-between and noother components may be referred to as such, in at least one example. Asyet another example, elements shown above/below one another, at oppositesides to one another, or to the left/right of one another may bereferred to as such, relative to one another. Further, as shown in thefigures, a topmost element or point of element may be referred to as a“top” of the component and a bottommost element or point of the elementmay be referred to as a “bottom” of the component, in at least oneexample. As used herein, top/bottom, upper/lower, above/below, may berelative to a vertical axis of the figures and used to describepositioning of elements of the figures relative to one another. As such,elements shown above other elements are positioned vertically above theother elements, in one example. As yet another example, shapes of theelements depicted within the figures may be referred to as having thoseshapes (e.g., such as being circular, straight, planar, curved, rounded,chamfered, angled, or the like). Further, elements shown intersectingone another may be referred to as intersecting elements or intersectingone another, in at least one example. Further still, an element shownwithin another element or shown outside of another element may bereferred as such, in one example.

Now turning to FIG. 3, graph 300 depicts example valve timings withrespect to a piston position, for an engine cylinder comprising 4valves: two intake valves and two exhaust valves, such as describedabove with reference to FIG. 1. The example of FIG. 3 is drawnsubstantially to scale, even though each and every point is not labeledwith numerical values. As such, relative differences in timings can beestimated by the drawing dimensions. However, other relative timings maybe used, if desired.

Continuing with FIG. 3, the cylinder is configured to receive intake viatwo intake valves and exhaust a first blowdown portion to a turbineinlet via a first exhaust valve (e.g., such as first, or blowdown,exhaust valves 8 shown in FIG. 1), exhaust a second scavenging portionto an intake passage via a second exhaust valve (e.g., such as second,or scavenge, exhaust valves 6 shown in FIG. 1) and non-combustedblowthrough air to the intake passage via the second exhaust valve. Byadjusting the timing of the opening and/or closing of the second exhaustvalve with that of the two intake valves, residual exhaust gases in thecylinder clearance volume may be cleaned out and recirculated as EGRalong with fresh intake blowthrough air.

Graph 300 illustrates an engine position along the x-axis in crank angledegrees (CAD). Curve 302 depicts piston positions (along the y-axis),with reference to their location from top dead center (TDC) and/orbottom dead center (BDC), and further with reference to their locationwithin the four strokes (intake, compression, power and exhaust) of anengine cycle.

During engine operation, each cylinder typically undergoes a four strokecycle including an intake stroke, compression stroke, expansion stroke,and exhaust stroke. During the intake stroke, generally, the exhaustvalves close and intake valves open. Air is introduced into the cylindervia the corresponding intake passage, and the cylinder piston moves tothe bottom of the cylinder so as to increase the volume within thecylinder. The position at which the piston is near the bottom of thecylinder and at the end of its stroke (e.g. when the combustion chamberis at its largest volume) is typically referred to by those of skill inthe art as bottom dead center (BDC). During the compression stroke, theintake valves and exhaust valves are closed. The piston moves toward thecylinder head so as to compress the air within combustion chamber. Thepoint at which the piston is at the end of its stroke and closest to thecylinder head (e.g. when the combustion chamber is at its smallestvolume) is typically referred to by those of skill in the art as topdead center (TDC). In a process herein referred to as injection, fuel isintroduced into the combustion chamber. In a process herein referred toas ignition, the injected fuel is ignited by known ignition means, suchas a spark plug, resulting in combustion. During the expansion stroke,the expanding gases push the piston back to BDC. A crankshaft convertsthis piston movement into a rotational torque of the rotary shaft.During the exhaust stroke, in a traditional design, exhaust valves areopened to release the residual combusted air-fuel mixture to thecorresponding exhaust passages and the piston returns to TDC. In thisdescription, the second exhaust (scavenge) valves may be opened afterthe beginning of the exhaust stroke and stay open until after the end ofthe exhaust stroke while the first exhaust (blowdown) valves are closedand the intake valves are opened to flush out residual exhaust gaseswith blowthrough air.

Curve 304 depicts a first intake valve timing, lift, and duration for afirst intake valve (Int_1) while curve 306 depicts a second intake valvetiming, lift, and duration for a second intake valve (Int_2) coupled tothe intake passage of the engine cylinder. Curve 308 depicts an exampleexhaust valve timing, lift, and duration for a first exhaust valve(Exh_1, which may correspond to first, or blowdown, exhaust valves 8shown in FIG. 1) coupled to a first exhaust manifold (e.g., blowdownexhaust manifold 84 shown in FIG. 1) of the engine cylinder, while curve310 depicts an example exhaust valve timing, lift, and duration for asecond exhaust valve (Exh_2, which may correspond to second, orscavenge, exhaust valves 6 shown in FIG. 1) coupled to a second exhaustmanifold (e.g., scavenge manifold 80 shown in FIG. 1) of the enginecylinder. As previously elaborated, the first exhaust manifold connectsa first exhaust valve to the inlet of a turbine in a turbocharger andthe second exhaust manifold connects a second exhaust valve to an intakepassage via an EGR passage. The first and second exhaust manifolds maybe separate from each other, as explained above.

In the depicted example, the first and second intake valves are fullyopened from a closed position at a common timing (curves 304 and 306),starting close to intake stroke TDC, just after CAD2 (e.g., at or justafter intake stroke TDC) and are closed after a subsequent compressionstroke has commenced past CAD3 (e.g., after BDC). Additionally, whenopened fully, the two intake valves may be opened with the same amountof valve lift L1 for the same duration of D1. In other examples, the twovalves may be operated with a different timing by adjusting the phasing,lift or duration based on engine conditions.

Now turning to the exhaust valves wherein the timing of the firstexhaust valve and the second exhaust valve is staggered relative to oneanother. Specifically, the first exhaust valve is opened from a closedposition at a first timing (curve 308) that is earlier in the enginecycle than the timing (curve 310) at which the second exhaust valve isopened from close. Specifically, the first timing for opening the firstexhaust valve is between TDC and BDC of the power stroke, before CAD1(e.g., before exhaust stroke BDC) while the timing for opening thesecond exhaust valve just after exhaust stroke BDC, after CAD1 butbefore CAD2. The first (curve 308) exhaust valve is closed before theend of the exhaust stroke and the second (curve 310) exhaust valve isclosed after the end of the exhaust stroke. Thus, the second exhaustvalve remains open to overlap slightly with opening of the intakevalves.

To elaborate, the first exhaust valve may be fully opened from closebefore the start of an exhaust stroke (e.g., between 90 and 40 degreesbefore BDC), maintained fully open through a first part of the exhauststroke and may be fully closed before the exhaust stroke ends (e.g.,between 50 and 0 degrees before TDC) to collect the blowdown portion ofthe exhaust pulse. The second exhaust valve (curve 310) may be fullyopened from a closed position just after the beginning of the exhauststroke (e.g., between 40 and 90 degrees past BDC), maintained openthrough a second portion of the exhaust stroke and may be fully closedafter the intake stroke begins (e.g., between 20 and 70 degrees afterTDC) to exhaust the scavenging portion of the exhaust. Additionally, thesecond exhaust valve and the intake valves, as shown in FIG. 3, may havea positive overlap phase (e.g., from between 20 degrees before TDC and40 degrees after TDC until between 40 and 90 degrees past TDC) to allowblowthrough with EGR. This cycle, wherein all four valves areoperational, may repeat itself based on engine operating conditions.

Additionally, the first exhaust valve may be opened at a first timingwith a first amount of valve lift L2 while the second exhaust valve maybe opened with a second amount of valve lift L3 (curve 310), where L3 issmaller than L2. Further still, the first exhaust valve may be opened atthe first timing for a duration D2 while the second exhaust valve may beopened for a duration D3, where D3 is smaller than D2. It will beappreciated that in alternate embodiments, the two exhaust valves mayhave the same amount of valve lift and/or same duration of opening whileopening at differently phased timings.

In this way, by using staggered valve timings, engine efficiency andpower can be increased by separating exhaust gases released at higherpressure (e.g., expanding blow-down exhaust gases in a cylinder) fromresidual exhaust gases at low pressure (e.g., exhaust gases that remainin the cylinder after blow-down) into the different passages. Byconveying low pressure residual exhaust gases as EGR along withblowthrough air to the compressor inlet (via the EGR passage and secondexhaust manifold), combustion chamber temperatures can be lowered,thereby reducing knock and spark retard from maximum torque. Further,since the exhaust gases at the end of the stroke are directed to eitherdownstream of a turbine or upstream of a compressor which are both atlower pressures, exhaust pumping losses can be minimized to improveengine efficiency.

Thus, exhaust gases can be used more efficiently than simply directingall the exhaust gas of a cylinder through a single, common exhaust portto a turbocharger turbine. As such, several advantages may be achieved.For example, the average exhaust gas pressure supplied to theturbocharger can be increased by separating and directing the blowdownpulse into the turbine inlet to improve turbocharger output.Additionally, fuel economy may be improved because blowthrough air isnot routed to the catalyst, being directed to the compressor inletinstead, and therefore, excess fuel may not be injected into the exhaustgases to maintain a stoichiometric ratio.

Exhaust gases that are recirculated to an engine intake through thescavenge manifold, flowing from the second exhaust gas valves during anexhaust stroke of a cylinder, as described above for FIG. 3, maycomprise a mixture of fresh air, burnt gas (e.g., combusted exhaustgas), and pushback fuel (e.g., unburnt fuel). Estimation of an EGRdilution rate, e.g. a fraction of burnt gas in an air mass at thecompressor inlet, arising from the recirculated exhaust gases, out ofthe combined exhaust gas, fresh blowthrough air, and pushback fuel, atthe engine intake may be achieved based on engine operating conditions.For example, engine speed and load may affect a combustion rate andhence the amount of exhaust gas generated. An amount of scavenge gas (amixture of fresh blowthrough air, bunt gas, and pushback fuel)recirculated from the scavenge manifold to the intake may be regulatedby a timing of an exhaust cam coupled to the scavenge exhaust valves.Additionally, the fraction of burnt gas recirculated to the enginecylinders may vary with a temperature of the scavenge gas.

A feedforward model may be used to approximate the EGR dilution forimproved performance of a split exhaust engine. The model determines aset of variables contributing to the dilution rate and includes: a totalEGR mass flow rate, a burnt gas mass flow rate, an EGR fuel mass, atemperature of EGR gases at an outlet of a venturi upstream of aturbocharger compressor, such as the venturi of ejector 56 of FIG. 1, achange in pressure across the venturi, a flow rate of a fresh air massentering the intake, and a flow rate of a fraction of burnt gas in theEGR gas at an inlet of the turbocharger compressor. Relationships andcontributions between the variables are illustrated in FIG. 4 in a flowdiagram 400.

Flow diagram 400 depicts an embodiment of a model used to estimate(e.g., calculate) an EGR rate of gases (containing exhaust gases,blowthrough air, and/or unburnt fuel) recirculated from the scavengeexhaust manifold to the intake passage. The EGR rate may also bereferred to as a burnt gas fraction and is based on contributions fromvariables representing mass flows, temperatures, and pressures, eachdetected from engine sensors, such as the sensors described above withrespect to FIGS. 1 and 2, arranged at certain locations of a splitexhaust engine, such as engine 10 of FIG. 1. At 402, the estimationmodel comprises calculating an air temperature of EGR gas, T_(EGR),flowing through a region upstream of a turbocharger in which a venturiupstream may be arranged, with reference to the venturi of ejector 56and compressor 162 of FIG. 1. For example, T_(EGR), may be thetemperature of the recirculated gases exiting the EGR passage andentering the intake passage, upstream of the compressor.

The T_(EGR) is the product of a temperature of EGR gas in a scavengemanifold, such as the scavenge manifold 80 of FIG. 1, and a temperaturecorrection accounting for loss of heat across the scavenge manifold. Thecalculation for the T_(EGR) may be described as,

T _(EGR) −T _(SM) *T _(EGRcorr)  (1)

where T_(SM) is the scavenge manifold temperature and T_(EGRcorr) is thetemperature correction. In one example, the scavenge manifoldtemperature may be a measured temperature obtained at a temperaturesensor positioned in the scavenge manifold, such as temperature sensor52 of FIG. 1, or from temperature sensors arranged in scavenge runners,such as the second exhaust runners 82 of FIG. 1, and averaged. Inanother example, T_(SM) may be determined based on engine speed andload. For example, T_(SM) may be mapped as a function of engine speedand load and plotted against brake mean effective pressure (BMEP), asshown in FIG. 5. The map may be pre-loaded in a memory of an enginecontroller, such as controller 12 of FIG. 1. Alternatively, look-uptables configured with data providing the T_(SM) according to a currentengine speed and load, during engine operation, may be stored in thecontroller's memory. In another example, the controller may beprogrammed to calculate the T_(SM from) a pre-set mathematical equationwhere current engine speed and engine load are inputs.

The temperature correction, T_(EGRcorr), accounts for heat transferacross a region between the location of the temperature sensor where theT_(SM) is measured and the region immediately upstream of the compressorinlet, if the temperature sensor is upstream of a scavenge EGR valve,such as the first EGR valve 54 of FIG. 1. As an example, differences inmeasured temperature along the scavenge manifold are shown in plot 500of FIG. 1 for one set of engine operating conditions.

Plot 500 of FIG. 5 shows an average scavenge runner temperature, an EGRcooler outlet temperature, EGRClrOut_Tmp (in embodiments of the splitexhaust engine which include an EGR cooler in the EGR passage), and atemperature of exhaust gas flowing out of scavenge manifold exhaustvalves (EGRVlvGasOut_Tmp) relative to BMEP (in psi or bar) along thex-axis. BMEP is defined as a theoretical average pressure that, ifuniformly imposed on pistons from the top to the bottom of each powerstroke, produces the measured power output of the engine. The averagescavenge runner temperature may be determined by temperaturemeasurements obtained from temperature sensors arranged in the scavengeexhaust runners. The EGRClrOut_Tmp may be measured by a temperaturesensor arranged downstream of the scavenge exhaust runners and upstreamof the scavenge EGR valve and the EGRVlvGasOut_Tmp may be measureddownstream of the scavenge exhaust valve and upstream of the venturi.The average scavenge runner temperature may be consistently higher thanboth the EGRClrOut_Tmp and the EGRVlvGasOut_Tmp, and the EGRClrOut_Tmpmay be consistently higher than the EGRVlvGasOut_Tmp. Estimateddifferences in temperature according to location of temperature sensorsmay be stored as look-up tables, such as plot 500 of FIG. 5, andreferred to by the controller.

In one example, the split exhaust engine may be configured withtemperature sensors at the scavenge exhaust runners but not downstreamof the scavenge EGR valve, e.g. the EGRVlvGasOut_Tmp may not be directlymeasured. However, the EGRVlvGasOut_Tmp provides the T_(SM) for theT_(EGR) calculation. The average scavenge runner temperature may be usedas an alternative and corrected based on the estimated difference intemperature according to plot 500. A magnitude of the T_(EGRcorr) mayvary depending on location of temperature sensors used for temperaturemeasurements and difference between ambient temperature and EGR gastemperature. Furthermore, vehicle speed may be included as an inputvariable that may affect the estimated T_(EGR). A manual or automatedcalibration effort may place the estimated steady state values equal tothe measured steady state values. The model may thereby be adjusted(i.e. calibrated) until the estimated data sufficiently agrees with themeasured data.

It will be appreciated that the description above for the calculation ofthe T_(EGR) is a non-limiting example of how the T_(EGR) may bedetermined and there may be numerous alternative methods to estimate theT_(EGR). In any example of the T_(EGR) calculation, however, thefeedforward, open loop model may provide the transient information andtemperature based data from the steady state model may provide thesteady state data. In the feedforward model, a bandpass filter may beapplied to each temperature signal that passes all signals greater than0.5 Hz to be added to data from the steady state model that may use aband pass filter that passes everything below 0.5 Hz. When addedtogether, a complete signal may be reconstructed. Data from a real (slowbut accurate) temperature measurement and a virtual temperaturemeasurement, estimated from other system parameters, is fused into asingle set of values.

Returning to FIG. 4, at 404 of the estimation model, the T_(EGR) is usedto determine the total EGR flow rate, EGR_(total). The EGR_(total) is atotal mass flow rate reaching the intake manifold and includes acombination of fresh air from the intake passage, introduced upstream ofthe venturi, and of the scavenge gas mixture comprising burnt gas, freshblowthrough air, and fuel vapor that is recirculated from the scavengemanifold to the intake passage, at a region upstream of the venturi. TheEGR_(total) may be described as:

EGR_(total) =f(Δ_(p) ,T _(EGR))  (3)

where the EGR_(total) is a function of a pressure differential, Δ_(p),across the venturi, or across the EGR valve and the venturi, upstream ofthe compressor at the region where the EGR passage couples to an intakepassage (e.g., intake passage 28 of FIG. 1), and the calculated T_(EGR).Determination of the Δ_(p) may be based on pressures measured bypressure sensors arranged upstream of the venturi in the intake passage,such as pressure sensor 51 of FIG. 1, and arranged upstream of theventuri and scavenge EGR valve in the EGR passage, such as pressuresensor 53 of FIG. 1.

In another example, the two flow rates (fresh air and recirculatedscavenge gas) flowing into the venturi and the flow rate exiting theventuri may be determine based on one flow (e.g., the fresh air) and theconcentration of EGR in the fresh air stream. Thus, the EGR flow ratemay be estimated based on measuring the air flow rate by a MAF sensor.

In one example, the pressure sensor in the EGR passage may be positionedupstream of the scavenge EGR valve (e.g., pressure sensor 53 and firstEGR valve 54 of FIG. 1) and the pressure measured may not account forchanges in pressure across the scavenge EGR valve. In another example,the pressure sensor may be positioned downstream of the scavenge EGRvalve, thereby detecting pressure after any potential change in pressuredue to the scavenge EGR valve. A pressure at an inlet of the venturi maybe approximated based on the pressure contributions from the intakepassage and the scavenge manifold, for example, by calculating aweighted average. A pressure downstream of the venturi, measured orinferred based on a calculated reduction in pressure provided by theventuri, is subtracted from the pressure at the inlet to give Δ_(p). Inanother example, the pressure downstream of the venturi, at thecompressor inlet may be assumed to be at atmospheric pressure, e.g., 1bar. A pressure differential across the venturi may be determined basedon the measured and calculated pressures, as a function of engine speedand load, and stored in the controller's memory. Alternatively, forimproved accuracy during low gas through the scavenge manifold and/orlow air flow through the intake passage, a differential pressure sensormay be used to measure a difference in pressure between an inlet and anoutlet of the venturi.

The Δ_(p) across the venturi may show significant variation in value.For example, the change in pressure across the venturi may range from 5kPa during low EGR flow rates up to 25 kPa during high EGR flow rates,the increase in EGR flow rate arising from increased engine speed andload. The change in Δ_(p) is mapped in plot 600 of FIG. 6A and plot 650of FIG. 6B, showing pressure relative to crank angle. A low EGR flowsituation is shown in FIG. 6A where little change is detected betweenthe pressure at the inlet of the venturi, averaged between pressure inthe intake passage (Intake) and pressure in the scavenge manifold andEGR passage (EGR (Avg)), and pressure downstream of the venturi(Compressor inlet). In FIG. 6B, a high EGR flow situation is shown andthe difference between the pressure upstream and downstream of theventuri (Δ_(p)) is much larger. The EGR_(total) may be used directly inthe calculation of the EGR rate at step 414.

At 406 of flow diagram 400, the mass flow rate of burnt gas (e.g.,combusted air/fuel or combusted exhaust gas), EGR_(bgas), is calculatedas a product of a mapped EGR burnt gas flow, which is a function ofengine speed and engine load, EGR_(bgas)′, and an EGR flow correction,which is a function of exhaust valve timing, EGR_(con)(Exh_vlv). Thecalculation may be represented by the following equation:

EGR_(bgas)=EGR_(bgas)′*EGR_(corr)  (4)

The EGR_(bgas)′ is a burned gas flow rate determined by mapping the rateas a function of engine speed and load. The EGR_(corr), is a correctionfactor applied to the mapped EGR_(bgas)′ that is calculated based on anactuation timing of the blowthrough and scavenge exhaust valves. TheEGR_(corr) accounts for an effect of exhaust valve cam timing on theEGR_(bgas). Maps of the EGR_(bgas)′ may be stored in the controller'smemory and consulted during engine operation based on the current enginespeed and load and corrected based on the current exhaust valve camtiming. The EGR_(bgas) may be applied to the determination of a rate ofair mass flow in the EGR gas at 410 of the estimation model.

At 410, the rate of air mass flow in the EGR gas, EGR_(am), may becalculated and used at 412 where a rate of fuel mass flow in the EGRgas, EGR_(fuel), may be calculated. Both calculations incorporate ablowthrough air/fuel ratio (AFR), BT_(afr), that may be determined at408. The BT_(afr) is a function of an amount of push back fuel and inone embodiment, may be assumed stoichiometric (e.g., 14.1:1 for agasoline engine). The BT_(afr) may be assumed stoichiometric and used tocalculate the EGR_(am) according to:

EGR_(am)=(EGR_(total)−EGR_(bgas))(BT_(afr)/(BT_(afr)+1))  (5)

The BT_(afr) may deviate from stoichiometric, however, as a result ofchanges in manifold absolute pressure (MAP) and/or scavenge manifoldpressure. This deviation may have little effect on the EGR_(am) but havea greater impact on the EGR_(fuel) calculation. Thus, a stoichiometricair/fuel ratio may be used to calculate the EGR_(am) while adjustment ofthe BT_(afr) in response to changes in engine speed, load and cam timingmay be desired to improve accuracy of the estimated EGR_(fuel).Alternatively, a UEGO may be used to estimate instantaneous fluidtemperature in the exhaust gas.

In another embodiment, the domination of blowthrough air over push-backfuel at higher MAP conditions may lead to a leaner BT_(afr). TheBT_(afr) may be mapped based on timing of the scavenge exhaust camversus the timing of the intake exhaust cam. The EGR_(fuel) may becalculated from the EGR_(am) and the BT_(afr) according to:

EGR_(fuel)=EGR_(am)(1/BT_(afr))  (6)

showing that the impact of the BT_(afr) on the EGR_(fuel) is much largerthan on the EGR_(am). To improve an accuracy of the estimation,determination of the change in BT_(afr) with a type of fuel injectionmay be desired. For example, different maps may be used depending on thetype of fuel injection. When fuel is introduced to the engine cylindersby port fuel injection (PFI), the AFR may be adjusted to compensate forpuddled fuel. This compensation may vary with engine speed and loadwhich affects a rate at which additional fuel from the puddled fuel iscombusted, e.g., the higher the engine speed and load, the fasterpuddled fuel is introduced through the intake valves which may bebalanced by adjusting the AFR to be leaner.

In another example, when the fuel is injected by DI, fuel is not puddledand since the blow through air is assumed stoichiometric and predominant(over pushback fuel) the AFR adjustment accounts for excess airexclusively. As another example, when a combination of PFI and DI isused for combustion, adjustment of the AFR to compensate for puddledfuel may be corrected based on the relative fractions of fuel injectedby PFI versus DI. In this way, maps or look-up tables for each of theinjection systems described above may be stored in the controllermemory, providing AFRs as functions of engine speed and load fordetermination of the BT_(afr).

At 414, the EGR rate is determined based on the mass air flow (MAF),measured by a MAF sensor such as MAF sensor 48 of FIG. 1, theEGR_(total), the EGR_(bgas), and the EGR_(fuel) according to:

EGR rate=EGR_(bgas)/(MAF+EGR_(total)−EGR_(fuel)  (7)

The EGR rate may be used to determine an engine dilution. Engineoperations such as spark advance and retard, fuel injection timing, andintake and exhaust cam timing may be adjusted in response to thecalculated engine dilution to improve a fuel efficiency and power outputof the engine. For example, less fuel may be injected during high EGRrates, or spark ignition may be advanced to achieve maximum torque.Overlap between opening of the blowdown exhaust valves and scavengeexhaust valves may be increased or decreased depending on the EGR rateto adjust a turbine speed or exhaust manifold pressure. Furthermore, ascavenge manifold bypass valve (SMBV), such as the SMBV 97 of FIG. 1,may be actuated to maintain a pressure in the scavenge manifold when theEGR rate increases.

The estimation of engine dilution by approximating the EGR rate viamapping of various parameters as described above for FIG. 4 may providea robust model for EGR flow during engine transients. However, there maybe events during which the EGR rate provided by the feedforward model ofFIG. 4 may deviate from the actual rate delivered to the engine intake.For example, engine transients leading to increases in EGR flow afterperiods of low EGR rate, such as during low engine speeds and loads, mayinclude a period of time for the newly generated scavenge gas to bemixed with intake air and ingested at the engine cylinders. In oneexample, this delay may occur over a period of six cylinder cycles.However, in other examples, the delay may occur faster or slower,depending on characteristics of the engine. During this interval, theEGR rate calculated according to the feedforward model of FIG. 4 may begreater than an actual EGR rate observed in the intake manifold butengine operations may be adjusted to the calculated rate. Engineefficiency and performance may be degraded as a result.

An offset of the estimated EGR rate from the actual rate during theperiod of time for recirculation of the scavenge gas mixture to theintake may be decreased by applying a temperature-based correction,ΔT_(corr), to the EGR rate, shown at 416 of FIG. 4. The ΔT_(corr) may bea steady state model that computes mass flow proportions of recirculatedscavenge gas to air, providing a concentration of recirculated gaswithin a main flow of a mixture of gases, e.g. the gas mixture flowinginto the venturi. The recirculated flow rate may be determined from themain flow, which is measured or inferred, and by subtracting the EGRflow rate, based on the feedforward model, from the recirculated flowrate, a flow of fuel vapor and air in the main flow may be estimated.Engine operations such as ignition timing, cam angles, etc., may beadjusted based on a calculated flow of fuel vapor and air when theengine is in a steady state. During engine transients, the feedforwardmodel may, in one example, be used to approximate EGR rate andcontinuously corrected based on the inferred steady state flow of fuelvapor and air, from temperature measurements, to accommodate delays inrecirculated gas delivery to the cylinders. In other examples, duringengine operations where the delay between generation of a burnt gasfraction in an air mass and delivery of the air mass to the intakemanifold is decreased, the feedforward model may be used exclusively toestimate the dilution rate.

A ratio of recirculated scavenge gas to air in a main air mass(including fuel vapor, fresh air, and combusted exhaust gas) to theintake of the split exhaust engine may be estimated using temperaturesmeasured at certain regions of the engine. For example, a temperature ofintake air entering the intake passage may be measured by a temperaturesensor arranged upstream of the venturi, such as the first intaketemperature sensor 3 of FIG. 1. A temperature of scavenge gasrecirculated through the scavenge manifold to the engine intake via theEGR passage may be measured by a temperature sensor positioned in theEGR passage, downstream of the scavenge EGR valve, such as temperaturesensor 52 of FIG. 1. A temperature of a combined gas mixture flowing outof the venturi outlet to the compressor inlet may be measured by atemperature sensor such as the second intake temperature sensor 31 ofFIG. 1. A list of variables used in the determination of circulated gasconcentration may include:

-   -   m_(air)=mass flow of intake air upstream of the venturi    -   m_(circ)=mass flow of gas mixture (e.g., burnt gas, fuel vapor,        fresh air) recirculated through scavenge manifold upstream of        the venturi    -   T_(air)=temperature of intake air upstream of the venturi    -   T_(circ)=temperature of recirculated gas mixture upstream of the        venturi    -   T_(comb)=temperature of combined intake air and recirculated gas        mixture downstream of the venturi and upstream of the compressor        inlet    -   Cp_(air)=specific heat of air    -   Cp_(circ)=specific heat of gas mixture recirculated through the        scavenge manifold to the intake

The ratio of recirculated gas to air is calculated as:

m _(cir) /m _(air)=[(T _(comb) −T _(air))/(T _(circ) −T _(air))]*(CP_(air) /CP _(circ))  (8)

The specific heats of air and circulated gas may be estimated by takinginto account the gaseous compositions of the air masses. For example,air may be assumed to comprise 21% oxygen, and 79% nitrogen. Using aspecific heat of oxygen, Cp_(oxy) and a specific heat of nitrogen,Cp_(nitro), the Cp_(air) may be calculated as:

Cp _(air)=(0.21*Cp _(oxy))+(0.79*Cp _(nitro))  (9)

where the values Cp_(oxy) and Cp_(nitro) are dependent on thetemperature of the air. In another example, the recirculated gas mixturemay comprise 21% water, 10.5% carbon dioxide, and 79% nitrogen (notethat the total is equal to 110.5% rather than 100%). The Cp_(circ) iscalculated according to:

Cp _(circ)=[(0.21*Cp _(water))+(0.105*Cp _(co2))+(0.79*Cp_(nitro))]*100/110.5  (10)

where the specific heats of water, Cp_(water), and of carbon dioxide,Cp_(co2), are also dependent on and vary according to the temperature ofthe recirculated gas mixture of the scavenge manifold.

In this way, the mass flow proportions of recirculated scavenge gas tofresh air may be calculated based on a principle of conservation ofenergy. The measured temperatures may deviate from actual temperatures,however, due to a relatively slow speed of heat conduction intemperature sensors. For example, a thermocouple or thermistor with a0.1 mm diameter may receive temperature information after a 0.05 seconddelay, adversely affecting an accuracy of the steady state model. Thespeed of temperature relay may be improved by using a smaller diameterthermocouple or thermistor with a faster response time.

For a split exhaust engine, such as engine 10 of FIG. 1, engineoperating parameters, such as spark advance, fuel injection, intakevalve cam timing, and exhaust valve cam timing, may be adjusted toenhance engine power output and performance based on the estimated EGRrate. The EGR rate is a dilution rate based on a mixture of gases,recirculated from a scavenge manifold to a region of the intake passageupstream of the compressor. The gas mixtures may include freshblowthrough air, fuel vapor (from unburnt fuel), and combusted exhaustgas (burnt gas) fractions of each type of gas may be determined from aplurality of engine variables and corrected by a temperature-basedsteady state model, as described in FIG. 4. An example routine forestimation of the EGR rate is depicted in a method 700 of FIG. 7 for thesplit exhaust engine. The EGR rate for the split exhaust engine may beapproximated during engine transients by the feedforward model. Thefeedforward model may be built upon measuring a pressure differentialacross the region of the intake passage upstream of the compressor inletand downstream of where the scavenge manifold couples to the intakepassage via an EGR passage, in which a flow constriction, such as aventuri, may be positioned. Mapped variables such as an exhaust camtiming, a scavenge manifold temperature, a burned gas flow rate, etc.may also be included in in the determination. The feedforward model maybe further improved by correction via the steady state model thatutilizes a calculated temperature differential to evaluateconcentrations of exhaust gas, fuel vapor and air in an air masscirculated through the split exhaust engine. Instructions for carryingout method 700 and the rest of the methods included herein may beexecuted by a controller based on instructions stored on a memory of thecontroller and in conjunction with signals received from sensors of theengine system, such as the sensors described above with reference toFIG. 1. The controller may employ engine actuators of the engine systemto adjust engine operation, according to the methods described below.

At 702, the method includes estimating and/or measuring the operatingconditions of the engine. These may include, for example, engine speedand load, MAP, a pressure within the scavenge manifold and within theEGR passage coupling the scavenge manifold to the intake air passage,oxygen content in gas flowing through an exhaust passage, a pressure ofthe intake air passage, an intake cam and an exhaust cam position, apressure at the compressor inlet, gas temperatures within the intakepassage and scavenge manifold, etc. For example, the exhaust campositioning may be determined from an exhaust valve position sensor,such as valve position sensor 157 of FIG. 2, and used to infer a timingof exhaust valve actuation. The controller may then use the exhaustvalve timing as a function of the engine speed and load to refer to acorresponding map or look-up table stored in the memory of thecontroller. The data provided by the map or look-up table may be used tocalculate a burnt gas contribution to estimate the EGR flow rate via thefeedforward model during transient engine conditions.

At 704 of the method, a first EGR rate may be estimated based on thesteady state model that is calculated from measurements obtained fromtemperature sensors in the intake passage, upstream of the venturi, inthe scavenge manifold or EGR passage, also upstream of the venturi, andfrom the intake passage between the venturi and the compressor inlet,e.g. at 416 of FIG. 4. The first EGR rate may be determined based ontemperatures measured from temperature sensors arranged in the intakepassage and EGR passage, upstream of where the EGR passage couples tothe intake passage, as well as downstream of where the EGR passagecouples to the intake passage. The calculated first EGR rate may becompared to a threshold flow rate at 706.

The threshold EGR flow rate may represent a rate that determines whenthe feedforward model (as explained above in reference to 414 and 416 inFIG. 4) for estimating a second EGR rate may be used. At flow ratesequal to or below the threshold, a pressure gradient across the venturimay be too low to be mapped for calculating total EGR mass flow rate forthe feedforward model. Thus, if the EGR rate is estimated to be at orbelow the threshold, the method returns to 707 to adjust engineoperations such as spark timing, intake and exhaust valve timings, fuelinjection etc., based on the first EGR rate.

If the first EGR rate is determined to be above the threshold, e.g., notat or below the threshold, the method continues to 708. In alternateembodiments, the method may proceed directly from 704 to 708 regardlessof the first EGR rate relative to the threshold.

At 708, a second EGR rate is calculated based on the mapping of engineparameters according to the feedforward model, as described above at402-414 in FIG. 4. At 710, the second EGR rate determined at 708 iscorrected (e.g., adjusted) based on the first EGR rate determined at704. For example, the controller may obtain temperature measurements tocompute the first EGR rate and apply the first EGR rate as a correctionfactor to the second EGR rate via a pre-set mathematical equation storedin a memory of the controller expressing a relationship between thefirst EGR rate and second EGR rate.

At 712, the method includes adjusting engine operation based on thecorrected EGR rate determined at 710. Adjusting engine operation mayinclude adjusting spark timing, exhaust valve timing (e.g., exhaustvalve cam timing), and/or fuel injection (e.g., adjusting an amount offuel injected by or a pulse width of one or more fuel injectors). Forexample, if the EGR rate is determined to increase as a result ofincreased engine speed and load, a spark timing may be advanced toaccount for a shorter period of time between spark and optimum peakpressure angle of the cylinders. As another example, the exhaust valvetiming may be modified to increase overlap between an opening ofblowdown exhaust valves, such as blowdown exhaust valves 8 of FIG. 1,and an opening of scavenge exhaust valves, such as exhaust valves 6 ofFIG. 1 to maintain an amount and pressure of scavenge gas recirculatedto the engine intake. Furthermore, a fuel injection timing may beadjusted according to spark advance to provide a desirable air-to-fuelratio at ignition.

Adjusting engine operations based on a combination of the feedforwardmodel and steady state model may have a significant impact onperformance of the split exhaust engine. For example, if the EGR ratewere estimated based on conventional EGR calculations configured with asingle set of exhaust valves and recirculating just burnt gas, thecalculated EGR rate may not account for introduction of blowthrough airinto the scavenge gas recirculated through the split exhaust engine. Theresulting AFR may be estimated to be too rich and lead to fueling to thecylinders that is too low. By using models that account for arecirculated mixture of gases, improved fueling and combustionefficiency may be achieved.

Example engine operation of a split exhaust engine, such as engine 10 ofFIG. 1, based on an estimate of an EGR rate (e.g., dilution rate) ofgases flowing from a scavenge manifold to an intake passage, upstream ofa compressor are now discussed with reference to FIG. 8. As elaboratedin map 800 of FIG. 8, engine load, determined by mass air flow into theengine intake, is shown at plot 801, a temperature of scavenge gas (e.g.blowthrough air, fuel vapor and burnt gas), measured within the scavengemanifold is shown at plot 802. A temperature of intake air, measured inthe intake passage upstream of a venturi arranged at a region where thescavenge manifold is coupled to the intake passage via an EGR passage,is shown at plot 804, and a temperature of a combined gas mixture,measured downstream of the venturi and upstream of a turbochargercompressor, is shown at plot 806. A position of a scavenge EGR valve,such as the scavenge EGR valve 54 of FIG. 1, is indicated at plot 808.EGR rates, representing dilution rates at the engine intake based on amixture of fresh intake air, blowthrough air, fuel vapor, and burnt gasare shown at plots 810, 812, 814, and 824. Plot 810 is an actual EGRrate and is compared with a steady state modeled EGR rate at plot 812,the steady state modeled EGR rate calculated based on a temperaturegradient across the venturi, such as at 416 of FIG. 4 and described byequations 8-10 above. An EGR rate determined from a feedforwardtransient model, as shown at 401-414 in FIG. 4, is depicted at plot 814.A position of a scavenge manifold bypass valve (SMBV), such as the SMBV97 of FIG. 1 is shown at plot 816. Adjustment of the positions of thescavenge EGR valve and SMBV may be based on EGR rates calculated fromthe steady state and feedforward models. Though the valve positions maybe shown as open and closed in FIG. 8, in alternate embodiments, thevalves may be adjusted into a plurality of positions between fully openand fully closed. A timing of intake valves is shown at plot 818 andexhaust valve timing is shown at plot 820 and a default timing isindicated by default line D1. In an embodiment where the scavengeexhaust valves and blowdown exhaust valves are controlled via a same camsystem, the exhaust valve timing at plot 820 may be the timing for boththe scavenge exhaust valves and the blowdown exhaust valves. A fuelingrate at the engine intake, e.g., fuel injection by DI, PFI or PFIDI, isshown at plot 822. The timing of intake and exhaust valves and fuelingrate may be adjusted based on a final, estimated EGR rate, shown at plot824, that may be determined from a combination of the steady state EGRrate shown at 812 and the feedforward EGR rate shown at 814 (or one orthe other). For example, at lower EGR rates, the estimated EGR rate at824 may be calculated solely from the steady state model. As the amountor flow rate of EGR increases, the feedforward model may provide anaccurate approximation of the actual EGR rate when corrected by thesteady state model.

Prior to t1 of map 800, engine load is relatively low (plot 801) and thetemperature of the scavenge gas (plot 802), is higher than thetemperatures of the intake air (plot 804) and combined gas mixture (plot806), the combined gas mixture comprising mostly intake air due to aclosed position of the scavenge EGR valve (plot 808). For example, thescavenge gas may be 90° C. while the intake air and combined gas mixturemay be at ambient temperature, such as 20° C. The actual EGR flow rate(plot 810), steady state modelled EGR flow rate (plot 812), andfeedforward modelled EGR flow rate (814) are at 0% of a maximumallowable EGR flow, also as a result of the closed position of thescavenge EGR valve. The estimated EGR rate (plot 824) is also at 0%. TheSMBV (plot 816) is also closed prior to t1. Intake and exhaust valvetimings are at the default timing (plots 818 and 820), and the fuelingrate (plot 822) is moderate, adjusted based on a stoichiometric AFRresulting from low engine load with no EGR flow from the scavengemanifold.

At t1, engine load begins to rise and combustion at the engine cylindersincreases so that exhaust gas generation also increases. As pressure inthe scavenge manifold accumulates, the scavenge EGR valve opens (orincreases an opening). The actual, measured EGR flow rises slightly to anonzero positive percentage between t1 and t2. At such low EGR flow, thefeedforward model may not detect an increase in the pressuredifferential across the venturi, thus the feedforward modeled EGR rateis estimated to remain at 0%. The estimated EGR rate is basedexclusively on the steady state model and also increases between t1 andt2. During this interval, the estimated EGR rate is not high enough toalter engine operations and the SMBV remains closed, intake valve andexhaust valve timings are at the default, and the fueling rate isunchanged.

At t2, engine load continues to increase, although at a slower rate.While the temperature of the intake air remains at ambient, the combinedgas mixture slowly rises in temperature due to a gradual increase intemperature of the scavenge gas. The scavenge EGR valve remains open,allowing an increase in flow of EGR to be delivered to the intake. TheEGR rate is sufficiently high to induce a measurable change in thepressure differential across the venturi, invoking an abrupt rise in thefeedforward modeled EGR flow rate while the calculation of the EGR rateby the steady state model shows a slight increase in EGR rate. The SMBVis maintained closed and intake valve and exhaust valve timings remainat the default timings. The estimated EGR rate converts to thefeedforward model and incorporates the correction based on the steadystate model. The fueling rate is decreased in response to the increasingestimated EGR rate of recirculated scavenge gas, containing fuel vaporand burnt gas, that is delivered to the engine cylinders (e.g., agreater fraction of exhaust gas is delivered to the engine cylinders andthus less fueling is desired).

At t3, engine load begins to plateau at elevated mass air flow and theactual EGR rate continues to increase briefly and then becomes level.The estimated EGR rate, based on the steady state-corrected feedforwardmodel, shows a similar levelling of the rate. In response to theestimated EGR rate, the fueling rate decreases between t3 to t4. Littlechange in the scavenge gas and the combined gas mixture temperaturesoccur between t2 and t3 but both temperatures increase between t3 andt4. The temperature of the intake air remains uniform and the SMBV ismaintained closed between t3 and t4. The intake valve timing and exhaustvalve timing are unchanged.

The actual EGR flow rate is constant between t3 and t4. However, thesteady state modeled EGR rate continues to rise due to increases in thescavenge and combined gas mixture temperatures. The feedforward modeledEGR rate also rises but in a linear manner due to an increase in thepressure differential across the venturi. The resulting estimated EGRrate gradually plateaus between t3 and t4.

At t4, the engine load remains high, and the estimated EGR rate, basedon the steady state-corrected feedforward model, reaches a maximum rateand plateaus during the interval between t4 and t5. The maximum rate maybe determined based on adjustment of the intake and exhaust valvetimings to provide a maximum amount of exhaust gas flow out of thescavenge exhaust valves. Shortly after t4, the intake valve timing isadvanced and the exhaust valve timing is retarded in response to theestimated EGR rate reaching the maximum rate. The portion of blowthroughair in the scavenge gas is increased, leading to a leaner AFR at theintake. In response to the change in the AFR, the fueling rate isincreased between t4 and t5 while the intake and exhaust valve timingsdeviate from the default timing. The SMBV is opened (or the opening isincreased relative to the closed position) at t4 to vent excess EGR toan exhaust passage. The scavenge gas and combined gas mixturetemperatures become relatively constant by t4 and show little changebetween t4 and t5.

At t5, engine load decreases but is still high enough that the scavengeEGR valve remains open. In response, the estimated EGR rate alsodecreases after t5, reflecting decreases in the steady state andfeedforward modeled EGR rates and simulating the actual EGR rates. TheSMBV is closed and valve timing returns to the default timing. Thefueling rate also decreases as the AFR approaches stoichiometric.

In this way, a dilution rate may be estimated for a split exhaust enginethat accounts for a unique configuration of intake and exhaust gas flowsof the engine. A feedforward model may be used during engine transientsthat determines the rate based on mapped engine parameters Suchparameters include a pressure differential across region downstream of amerging point between an intake passage and an EGR passage coupled to ascavenge manifold and upstream of a turbocharger compressor, as well asexhaust valve timing as a function of engine speed and load, and ablowthrough air-to-fuel ratio. The feedforward model may be supported bya steady state model that estimates a burnt gas fraction of the gasmixture delivered to the engine intake based on a temperaturedifferential across the region downstream of the merging point betweenthe intake passage and the EGR passage. The steady state model may beused at low EGR rates as an alternative to the feedforward model andalso used as a correction factor for the feedforward model to accountfor delays between generation of a particular burnt gas fraction andtransport of the gas mixture to the combustion chambers. By providing arobust method for estimating the EGR rate, timing of engine operationsmay be adjusted accordingly to increase engine performance. Thetechnical effect of estimating the EGR rate by the combination of thesteady state model and feedforward model is to more accurately estimatethe EGR rate of recirculated gas from a scavenge manifold containingeach of fresh, blowthrough air, combusted exhaust gas, and unburnt fuel.As a result of more accurately estimating the EGR rate, engine operatingparameters (such as spark timing, fuel injection, and cylinder valvetimings) may be more accurately adjusted to increase engine power outputand reduce a likelihood of engine knock.

Note that the example control and estimation routines included hereincan be used with various engine and/or vehicle system configurations.The control methods and routines disclosed herein may be stored asexecutable instructions in non-transitory memory and may be carried outby the control system including the controller in combination with thevarious sensors, actuators, and other engine hardware. The specificroutines described herein may represent one or more of any number ofprocessing strategies such as event-driven, interrupt-driven,multi-tasking, multi-threading, and the like. As such, various actions,operations, and/or functions illustrated may be performed in thesequence illustrated, in parallel, or in some cases omitted. Likewise,the order of processing is not necessarily required to achieve thefeatures and advantages of the example embodiments described herein, butis provided for ease of illustration and description. One or more of theillustrated actions, operations and/or functions may be repeatedlyperformed depending on the particular strategy being used. Further, thedescribed actions, operations and/or functions may graphically representcode to be programmed into non-transitory memory of the computerreadable storage medium in the engine control system, where thedescribed actions are carried out by executing the instructions in asystem including the various engine hardware components in combinationwith the electronic controller.

It will be appreciated that the configurations and routines disclosedherein are exemplary in nature, and that these specific embodiments arenot to be considered in a limiting sense, because numerous variationsare possible. For example, the above technology can be applied to V-6,1-4, 1-6, V-12, opposed 4, and other engine types. The subject matter ofthe present disclosure includes all novel and non-obvious combinationsand sub-combinations of the various systems and configurations, andother features, functions, and/or properties disclosed herein.

As one embodiment, a method includes determining a dilution rate of gasrecirculated from a first set of exhaust valves to an intake passage viaa recirculation passage based on a temperature of gases in each of therecirculation passage and the intake passage, upstream and downstream ofwhere the EGR passage couples to the intake passage, while flowingexhaust gas from a second set of exhaust valves to a turbochargerturbine and not to the intake passage. In a first example of the method,each cylinder includes one valve from each of the first and second setof exhaust valves. A second example of the method optionally includesthe first example, and further includes adjusting one or more of fuelinjection to engine cylinders and spark timing based on the determineddilution rate. A third example of the method optionally includes one ormore of the first and second examples, and further includes, whereindetermining the dilution rate of gas recirculated from the first set ofexhaust valves to the intake passage includes determining the dilutionrate of gas recirculated from the first set of exhaust valves to theintake passage, upstream of a turbocharger compressor based on each of afirst temperature of intake air in the intake passage, upstream of theturbocharger compressor, a second temperature of recirculated gases inthe recirculation passage, upstream of where the recirculation passagecouples to the intake passage, and a third temperature of combined airand recirculated gasses in the intake passage, downstream of theturbocharger compressor. A fourth example of the method optionallyincludes one or more of the first through third examples, and furtherincludes, wherein determining the dilution rate of gas recirculated fromthe first set of exhaust valves to the intake passage includesdetermining a first dilution rate of gas recirculated from the first setof exhaust valves to the intake passage based on each of engine speed,engine load, a pressure differential across a flow constriction arrangedin the intake passage downstream of where the recirculation passagecouples to the intake passage, and a cam timing of the first set ofexhaust valves, determining a second dilution rate of gas recirculatedfrom the first set of exhaust valves to the intake passage based on thetemperature of gases in each of the recirculation passage and the intakepassage, upstream and downstream of where the recirculation passagecouples to the intake passage, and correcting the first dilution ratewith the second dilution rate to determine a final, corrected dilutionrate. A fifth example of the method optionally includes one or more ofthe first through fourth examples, and further includes, adjustingengine operation based on the final, corrected dilution rate. A sixthexample of the method optionally includes one or more of the firstthrough fifth examples, and further includes, opening the first set ofexhaust valves at a different timing in an engine cycle than the secondset of exhaust valves. A seventh example of the method optionallyincludes one or more of the first through sixth examples, and furtherincludes, wherein the gas recirculated from the first set of exhaustvalves includes a portion of each of burnt combustion gases, freshblowthrough air, and unburnt fuel. An eighth example of the methodoptionally includes one or more of the first through seventh examples,and further includes, combusting air and fuel in each engine cylinderand then: first, flowing a first portion of combusted gases to theturbocharger turbine disposed in an exhaust passage via the second setof exhaust valves; second, flowing a second portion of combusted gasesto the intake passage via the first set of exhaust valves; and, third,flowing fresh, blowthrough air to the intake passage via the first setof exhaust valves. A ninth example of the method optionally includes oneor more of the first through eighth examples, and further includes, notflowing blowthrough air to the turbocharger compressor and furthercomprising flowing a portion of the second portion of combusted gases tothe exhaust passage, downstream of the turbocharger turbine.

In another embodiment, a method includes flowing gases from a first setof exhaust valves to a compressor disposed in an intake passage andflowing exhaust gas from a second set of exhaust valves to a turbinedisposed in an exhaust passage and not to the intake passage, where eachcylinder of a plurality of engine cylinders includes one valve of thefirst set of exhaust valves and one valve of the second set of exhaustvalves, determining a dilution rate of the gases flowing from the firstset of exhaust valves to the compressor based on a temperature of eachof: the gases flowing from the first set of exhaust valves to thecompressor, before the gases enter the intake passage, intake airflowing in the intake passage, upstream of where the gases from thefirst set of exhaust valves enter the intake passage, and combined gasesflowing through the intake passage, downstream of where the gases fromthe first set of exhaust valves enter the intake passage, and adjustingone or more of fuel injection to the plurality of engine cylinders andspark timing based on the determined dilution rate. In a first exampleof the method, the gases from the first set of exhaust valves includes acombination of combusted exhaust gases and fresh, blowthrough air, wherean amount of fresh, blowthrough air is based on a valve opening overlapperiod between the first set of exhaust valves and intake valves of eachcylinder, and wherein the combusted exhaust gases from the second set ofexhaust valves does not contain fresh, blowthrough air. A second exampleof the method optionally includes the first example and further includeswherein the determined dilution rate of the gases flowing from the firstset of exhaust valves to the compressor is a first determined dilutionrate, further comprising determining a second dilution rate of gasesflowing from the first set of exhaust valves to the compressor based oneach of engine speed, engine load, a pressure differential across a flowconstriction arranged in the intake passage, and a cam timing of thefirst set of exhaust valves, and further comprising correcting thesecond determined dilution rate with the first determined dilution rateto determine a final dilution rate and adjusting one or more of fuelinjection to the plurality of engine cylinders and spark timing based onthe final determined dilution rate. A third example of the methodoptionally includes one or more of the first and second examples, andfurther includes, wherein flowing gases from the first set of exhaustvalves to the compressor includes flowing a combination of combustedexhaust gases, fresh blow through air, and unburnt fuel from the firstset of exhaust valves to the intake passage, upstream of the compressor,via an exhaust gas recirculation (EGR) passage including an EGR valveand wherein the second determined dilution rate is further based on adifferential pressure across the EGR valve and the flow constriction. Afourth example of the method optionally includes one or more of thefirst through third examples, and further includes, wherein adjustingone or more of fuel injection to the plurality of engine cylinders andspark timing based on the determined dilution rate includes actuatingone or more fuel injectors to increase an amount of fuel injected intothe plurality of engine cylinders in response to the determined dilutionrate decreasing.

As another embodiment, a system for an engine includes a first set ofexhaust valves exclusively coupled to a first exhaust manifold, thefirst exhaust manifold coupled to an intake passage, upstream of aturbocharger compressor, via an exhaust gas recirculation (EGR) passage,the EGR passage including an EGR valve, a second set of exhaust valvesexclusively coupled to a second exhaust manifold coupled to an exhaustpassage, upstream of a turbocharger turbine disposed in the exhaustpassage, a plurality of engine cylinders, each including one of thefirst set of exhaust valves and one of the second set of exhaust valves,and a controller including memory with instructions stored thereon fordetermining a dilution rate of gases flowing from the first set ofexhaust valves to the turbocharger compressor via the EGR passage basedon a first temperature of gases in the EGR passage, a second temperatureof gases in the intake passage, upstream of where the EGR passagecouples to the intake passage, and a third temperature of gases in theintake passage, downstream of the turbocharger compressor, and adjustingspark timing and fuel injection to the plurality of engine cylindersbased on the determined dilution rate. In a first example of the system,the first set of exhaust valves open at a different timing than thesecond set of exhaust valves and wherein there is a valve overlap periodbetween the first set of exhaust valves and intake valves of theplurality of engine cylinders where the one exhaust valve and intakevalves of each cylinder are both open while the second set of exhaustvalves are closed. A second example of the system optionally includesthe first example, and further includes a bypass passaged coupledbetween the first exhaust manifold and the exhaust passage, downstreamof the turbocharger turbine. A third example of the system optionallyincludes one or more of the first and second examples, and furtherincludes, a first temperature sensor positioned in the EGR passage, asecond temperature sensor positioned in the intake passage, upstream ofwhere the EGR passage couples to the intake passage, and a thirdtemperature sensor positioned in the intake passage, downstream of theturbocharger compressor and wherein the first temperature, secondtemperature, and third temperature are measured temperatures determinedfrom outputs of each of the first temperature sensor, second temperaturesensor, and third temperature sensor, respectively. A fourth example ofthe system optionally includes one or more of the first through thirdexamples, and further includes, wherein the instructions further includeinstructions for determining the dilution rate based on engine speed,engine load, a cam timing of the first set of exhaust valves, and adifferential pressure across the EGR valve and a flow constrictionpositioned in the intake passage downstream of where the EGR passagecouples to the intake passage.

The following claims particularly point out certain combinations andsub-combinations regarded as novel and non-obvious. These claims mayrefer to “an” element or “a first” element or the equivalent thereof.Such claims should be understood to include incorporation of one or moresuch elements, neither requiring nor excluding two or more suchelements. Other combinations and sub-combinations of the disclosedfeatures, functions, elements, and/or properties may be claimed throughamendment of the present claims or through presentation of new claims inthis or a related application. Such claims, whether broader, narrower,equal, or different in scope to the original claims, also are regardedas included within the subject matter of the present disclosure.

1. A method comprising: determining a dilution rate of gas recirculatedfrom a first set of exhaust valves to an intake passage via arecirculation passage based on a temperature of gases in each of therecirculation passage and the intake passage, upstream and downstream ofwhere the EGR passage couples to the intake passage, while flowingexhaust gas from a second set of exhaust valves to a turbochargerturbine and not to the intake passage.
 2. The method of claim 1, whereineach cylinder includes one valve from each of the first and second setof exhaust valves.
 3. The method of claim 1, further comprisingadjusting one or more of fuel injection to engine cylinders and sparktiming based on the determined dilution rate.
 4. The method of claim 1,wherein determining the dilution rate of gas recirculated from the firstset of exhaust valves to the intake passage includes determining thedilution rate of gas recirculated from the first set of exhaust valvesto the intake passage, upstream of a turbocharger compressor based oneach of a first temperature of intake air in the intake passage,upstream of the turbocharger compressor, a second temperature ofrecirculated gases in the recirculation passage, upstream of where therecirculation passage couples to the intake passage, and a thirdtemperature of combined air and recirculated gasses in the intakepassage, downstream of the turbocharger compressor.
 5. The method ofclaim 1, wherein determining the dilution rate of gas recirculated fromthe first set of exhaust valves to the intake passage includesdetermining a first dilution rate of gas recirculated from the first setof exhaust valves to the intake passage based on each of engine speed,engine load, a pressure differential across a flow constriction arrangedin the intake passage downstream of where the recirculation passagecouples to the intake passage, and a cam timing of the first set ofexhaust valves, determining a second dilution rate of gas recirculatedfrom the first set of exhaust valves to the intake passage based on thetemperature of gases in each of the recirculation passage and the intakepassage, upstream and downstream of where the recirculation passagecouples to the intake passage, and correcting the first dilution ratewith the second dilution rate to determine a final, corrected dilutionrate.
 6. The method of claim 5, further comprising adjusting engineoperation based on the final, corrected dilution rate.
 7. The method ofclaim 1, further comprising opening the first set of exhaust valves at adifferent timing in an engine cycle than the second set of exhaustvalves.
 8. The method of claim 1, wherein the gas recirculated from thefirst set of exhaust valves includes a portion of each of burntcombustion gases, fresh blowthrough air, and unburnt fuel.
 9. The methodof claim 1, further comprising combusting air and fuel in each enginecylinder and then: first, flowing a first portion of combusted gases tothe turbocharger turbine disposed in an exhaust passage via the secondset of exhaust valves; second, flowing a second portion of combustedgases to the intake passage via the first set of exhaust valves; and,third, flowing fresh, blowthrough air to the intake passage via thefirst set of exhaust valves.
 10. The method of claim 9, furthercomprising not flowing blowthrough air to the turbocharger compressorand further comprising flowing a portion of the second portion ofcombusted gases to the exhaust passage, downstream of the turbochargerturbine.
 11. A method, comprising: flowing gases from a first set ofexhaust valves to a compressor disposed in an intake passage and flowingexhaust gas from a second set of exhaust valves to a turbine disposed inan exhaust passage and not to the intake passage, where each cylinder ofa plurality of engine cylinders includes one valve of the first set ofexhaust valves and one valve of the second set of exhaust valves;determining a dilution rate of the gases flowing from the first set ofexhaust valves to the compressor based on a temperature of each of: thegases flowing from the first set of exhaust valves to the compressor,before the gases enter the intake passage, intake air flowing in theintake passage, upstream of where the gases from the first set ofexhaust valves enter the intake passage, and combined gases flowingthrough the intake passage, downstream of where the gases from the firstset of exhaust valves enter the intake passage; and adjusting one ormore of fuel injection to the plurality of engine cylinders and sparktiming based on the determined dilution rate.
 12. The method of claim11, wherein the gases from the first set of exhaust valves includes acombination of combusted exhaust gases and fresh, blowthrough air, wherean amount of fresh, blowthrough air is based on a valve opening overlapperiod between the first set of exhaust valves and intake valves of eachcylinder, and wherein the combusted exhaust gases from the second set ofexhaust valves does not contain fresh, blowthrough air.
 13. The methodof claim 11, wherein the determined dilution rate of the gases flowingfrom the first set of exhaust valves to the compressor is a firstdetermined dilution rate, further comprising determining a seconddilution rate of gases flowing from the first set of exhaust valves tothe compressor based on each of engine speed, engine load, a pressuredifferential across a flow constriction arranged in the intake passage,and a cam timing of the first set of exhaust valves, and furthercomprising correcting the second determined dilution rate with the firstdetermined dilution rate to determine a final dilution rate andadjusting one or more of fuel injection to the plurality of enginecylinders and spark timing based on the final determined dilution rate.14. The method of claim 13, wherein flowing gases from the first set ofexhaust valves to the compressor includes flowing a combination ofcombusted exhaust gases, fresh blow through air, and unburnt fuel fromthe first set of exhaust valves to the intake passage, upstream of thecompressor, via an exhaust gas recirculation (EGR) passage including anEGR valve and wherein the second determined dilution rate is furtherbased on a differential pressure across the EGR valve and the flowconstriction.
 15. The method of claim 13, wherein adjusting one or moreof fuel injection to the plurality of engine cylinders and spark timingbased on the determined dilution rate includes actuating one or morefuel injectors to increase an amount of fuel injected into the pluralityof engine cylinders in response to the determined dilution ratedecreasing.
 16. A system for an engine, comprising: a first set ofexhaust valves exclusively coupled to a first exhaust manifold, thefirst exhaust manifold coupled to an intake passage, upstream of aturbocharger compressor, via an exhaust gas recirculation (EGR) passage,the EGR passage including an EGR valve; a second set of exhaust valvesexclusively coupled to a second exhaust manifold coupled to an exhaustpassage, upstream of a turbocharger turbine disposed in the exhaustpassage; a plurality of engine cylinders, each including one of thefirst set of exhaust valves and one of the second set of exhaust valves;and a controller including memory with instructions stored thereon for:determining a dilution rate of gases flowing from the first set ofexhaust valves to the turbocharger compressor via the EGR passage basedon a first temperature of gases in the EGR passage, a second temperatureof gases in the intake passage, upstream of where the EGR passagecouples to the intake passage, and a third temperature of gases in theintake passage, downstream of the turbocharger compressor; and adjustingspark timing and fuel injection to the plurality of engine cylindersbased on the determined dilution rate.
 17. The system of claim 16,wherein the first set of exhaust valves open at a different timing thanthe second set of exhaust valves and wherein there is a valve overlapperiod between the first set of exhaust valves and intake valves of theplurality of engine cylinders where the one exhaust valve and intakevalves of each cylinder are both open while the second set of exhaustvalves are closed.
 18. The system of claim 16, further comprising abypass passaged coupled between the first exhaust manifold and theexhaust passage, downstream of the turbocharger turbine.
 19. The systemof claim 16, further comprising a first temperature sensor positioned inthe EGR passage, a second temperature sensor positioned in the intakepassage, upstream of where the EGR passage couples to the intakepassage, and a third temperature sensor positioned in the intakepassage, downstream of the turbocharger compressor and wherein the firsttemperature, second temperature, and third temperature are measuredtemperatures determined from outputs of each of the first temperaturesensor, second temperature sensor, and third temperature sensor,respectively.
 20. The system of claim 16, wherein the instructionsfurther include instructions for determining the dilution rate based onengine speed, engine load, a cam timing of the first set of exhaustvalves, and a differential pressure across the EGR valve and a flowconstriction positioned in the intake passage downstream of where theEGR passage couples to the intake passage.